Design and Manufacturing of Loop Heat Pipes for ...

Design and Manufacturing of Loop Heat Pipes for ElectronicsCooling

by

Shane Storring

A thesis submitted to the Faculty of Graduate Studies and Research

in partial fulfilment of the degree requirements of

Master’s of Applied Science

In Aerospace Engineering

Ottawa-Carleton Institute for

Mechanical and Aerospace Engineering

Department of Mechanical and Aerospace Engineering

Carleton University

Ottawa, Ontario, Canada

February, 2006

Copyright ©

2006 - Shane Storring

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Abstract

A loop heat pipe (LHP) is a two-phase heat transfer device that uses the latent heat of

vaporization of a working fluid for heat transfer. Due to their inherent robustness, LHPs

are ideal candidates to meet the ever increasing thermal challenges facing the electronics

and spacecraft industries. Through miniaturization, LHPs have the potential to efficiently

transport, spread, and dissipate heat in advanced electronics packages where the heat

dissipation is rapidly increasing with decreasing volume.

In this study, three LHPs were designed, manufactured, and tested to help gain the

knowledge necessary for developing smaller and more efficient designs for use in future

Canadian space missions. A detailed development process was discussed and included

initial design criteria and selection, material preparation and treatment, assembly,

evacuation, charging, and sealing. A series of tests were conducted to investigate the

thermal performance characteristics of each unit. Tests were performed under ambient

conditions to study the effect of fluid inventory on LHP start-up, steady-state operating

temperature, effective thermal resistance, and overall heat transfer coefficient. Alternate

configurations were investigated to measure the effect of orientation and periodic heating

on thermal performance. LHP operation was also examined for temperature hysteresis

during rapid power variations.

iii


Numerical models were developed to aid in the design phase of the study by

estimating the properties of various working fluids, calculating fluid inventory, and sizing

the compensation chamber. The models also provide the foundation for future work on a

1-D steady state solver to predict operational characteristics such as steady-state

temperatures, system pressure drop and mass flow rate for any given design.

iv


Acknowledgements

First of all, I would like to express my sincere gratitude and appreciation to my thesis

supervisor, Dr. Tarik Kaya, for giving me this opportunity and for his help and guidance

throughout the project. I would also like to acknowledge his flexibility in allowing me to

pursue my goals and career outside of the University. I would like to acknowledge Feng

Cai and Michel Garcia for their many hours of help in the lab as well as the complete

staff in the machine shop for their many lessons on how not to injure myself. I would

like to give credit to everyone who helped review my thesis, their time is very much

appreciated. Also, I would like to recognize the staff at Bristol Aerospace their flexibility

and support during the final phase of my thesis. Finally, I would like to thank my family

and friends for their love and support during the most difficult times. They are always

there for me when I need it the most and give me the confidence I need as I embark on a

new journey in life.

v


Table of Contents

Abstract................................................................................................................................... iii

Acknowledgements................................................................................................................. v

Table of Contents.................................................................................................................. vi

List of Tables............................................................................................................................x

List of Figures........................................................................................................................ xi

Nomenclature .................................................................................................................... xiv

Chapter 1: Introduction............................. 1

1.1 Background.................................................................................................................... 3

1.1.1 Conventional Heat Pipes........................................................................................ 3

1.1.2 Capillary Pumped Loops (CPL).............................................................................5

1.1.3 Loop Heat Pipes (LHP).......................................................................................... 6

1.1.4 Miniature Loop Heat Pipe (MLHP)...................................................................... 7

1.2 Research Objectives.......................................................................................................9

1.3 Organization................................................................................................................. 10

Chapter 2: Literature Review............................................................................................ 12

2.1 LHP Operating Principles........................................................................................... 12

2.1.1 Thermodynamic Cycle......................................................................................... 14

2.2 LHP Operating Characteristics.................................................................................... 18

2.2.1 Loop Operating Temperature...............................................................................18

vi


2.2.2 Start-Up................................................................................................................. 23

2.2.3 Temperature Hysteresis........................................................................................ 26

2.2.4 Effect of Orientation............................................................................................. 27

2.2.5 Effect of Non-Condensable Gases...................................................................... 29

2.3 Design of Loop Heat P ipes....................................................................................... 29

2.3.1 Working Fluids......................................................................................................33

2.3.2 Primary and Secondary Wicks.............................................................................36

2.3.3 Component Sizing.................................................................................................38

2.3.4 Fluid Inventory and Compensation Chamber.................................................... 38

Chapter 3: LHP Design........................................................................................................44

3.1 Problem Statement.......................................................................................................44

3.2 Design Approach..........................................................................................................45

3.3 Wick Material Selection.............................................................................................. 45

3.4 Working Fluid Selection............................................................................................. 47

3.5 Case Material Selection............................................................................................... 49

3.6 Loop Sealing................................................................................................................ 50

3.7 Component Sizing........................................................................................................51

3.8 Wick Structure Design.................................................................................................52

3.9 Mathematical Modeling.............................................................................................. 54

3.10 Final LHP Design and Layout...................................................................................59

Chapter 4: LHP Manufacturing........................................................................................ 61

4.1 Fabrication.................................................................................................................... 61

4.1.1 Case Materials.......................................................................................................62

4.1.2 W ick...................................................................................................................... 64

4.1.3 Heat Source and Heat Sink Saddles.................................................................... 66

4.2 Component Cleaning....................................................................................................69

4.3 Assembly Procedure.....................................................................................................70

4.3.1 Swagelok Fittings................................................................................................72

vii


4.4 Evacuation and Charging Procedures.........................................................................75

4.5 LHP Sealing................................................................................................................. 79

Chapter 5: LHP Testing...................................................................................................... 81

5.1 Experimental Approach................................................................................................81

5.2 Test Specimen and Experimental Setup..................................................................... 81

5.2.1 Orientation of Test U nit....................................................................................... 83

5.2.2 Instrumentation......................................................................................................83

5.2.3 Insulation and Test Frame.................................................................................... 85

5.2.4 Evaporator-Heat Source Assembly..................................................................... 86

5.2.5 Sink Assembly.......................................................................................................87

5.3 Operating Procedure.....................................................................................................90

5.4 Test Phase 1: Preliminary Testing...............................................................................92

5.4.1 Baseline Tests........................................................................................................92

5.4.2 Test Apparatus #1..................................................................................................94

5.4.3 Test Apparatus #2..................................................................................................98

5.4.4 Test Apparatus #3................................................................................................100

5.5 Test Phase 2: Experimental Study of the Effect of Fluid Inventory on LHP

Performance......................................................................................................................101

5.5.1 Fluid Inventory....................................................................................................101

5.5.2 Start-Up............................................................................................................... 102

5.5.3 Steady-State Operating Temperatures............................................................... I l l

5.5.4 Effective Thermal Resistance.............................................................................116

5.5.5 Heat Transfer Coefficient................................................................................... 118

5.6 Test Phase 3: Experimental Study of LHP Operating Characteristics................... 120

5.6.1 Effect of Elevation................................................................................... 120

5.6.2 Periodic Heating..................................................................................................122

5.6.3 Temperature Hysteresis...................................................................................... 123

Chapter 6: Mathematical Modeling................................................................................126

viii


6.1 Software Development...............................................................................................126

6.1.1 Graphical User Interface (GUI) & Input/Output Parameters...........................126

6.1.2 Fluid Properties...................................................................................................129

6.1.3 Figure of M erit.................................................................................................... 130

6.1.4 Fluid Inventory.................................................................................................... 131

Chapter 7: Conclusions and Recommendations............................................................ 133

7.1 Summary and Conclusions.........................................................................................133

7.2 Recommendations...................................................................................................... 137

References............................................................................................................................ 141

Appendix A .......................................................................................................................... 143

ix


List of Tables

Table 2-1: Tested LHP Material-Working Fluid Combinations......................................... 31

Table 2-2: Components and their Influence on Design Requirements.............................. 32

Table 2-3: Operating Temperature Range of Various Fluids.............................................. 34

Table 3-1: Potential Working Fluids and their Properties.................................................. 48

Table 3-2: Detailed Properties of each Test Unit................................................................. 59

Table 5-1: Summary of Fluid Charges................................................................................ 102

Table 5-2: Summary of Start-up Results for a Fluid Charge of 20 g ................................108

Table 5-3: Summary of Start-up Tim es.............................................................................. I l l

x


List of Figures

Figure 1-1: LHP Configuration................................................................................................2

Figure 1-2: Conventional Heat Pipe........................................................................................ 4

Figure 2-1: Pressure-Temperature Curve of Steady-State LHP Operation.........................15

Figure 2-2: Simplified Thermal Network of the LHP M odel............................................. 19

Figure 2-3: Variable and Fixed Conductance Modes...........................................................23

Figure 2-4: LHP Start-up Scenarios in the Evaporator........................................................25

Figure 2-5: Typical Temperature Hysteresis Trend.............................................................26

Figure 2-6: Loop Heat Pipe Design Procedure.................................................................... 30

Figure 2-7: Figure of Merit for Various Working Fluids.....................................................36

Figure 2-8: Fluid Distribution Scheme 1...............................................................................41

Figure 3-1: Design of Evaporator and Primary Wick Interface.......................................... 54

Figure 3-2: Solid Model of Wick Structure with Vapour Grooves.....................................54

Figure 3-3: EASY 2000 User Interface and Input Parameters............................................ 56

Figure 3-4: Calculation of Fluid Inventory, Reservoir Volume, and LHP Mass...............57

Figure 3-5: Pressure Balance and Maximum Heat Flux Estimate for Given Design 58

Figure 4-1: Loop Heat Pipe Manufacturing Process............................................................62

Figure 4-2: Band Saw used to Cut Stainless Steel Tubing.................................................. 63

Figure 4-3: Lathe used to Round-Off Edges of Tubing.......................................................63

Figure 4-4: Lathe used for Diameter and Four Circumferential Grooves.......................... 65

Figure 4-5: Mill used to Machine Four Wick Axial Grooves............................................. 66

Figure 4-6: Wick Plastic Cap.................................................................................................66

Figure 4-7: Milling Machine and Drill Press used for Manufacturing Saddles.................67

Figure 4-8: Threads Machined using a 10-32 Hole Tap......................................................67

Figure 4-9: Final Evaporator Saddle Assembly................................................................... 68

xi


Figure 4-10: Final Condenser and Subcooler Saddle Assembly......................................... 68

Figure 4-11: Tube-in-Tube Parallel Flow Heat Exchanger................................................. 69

Figure 4-12: Wick-Evaporator Insertion...............................................................................71

Figure 4-13: Evaporator Wall-Wick Contact after Insertion.............................................. 71

Figure 4-14: Swagelok Two-Ferrule Design.........................................................................72

Figure 4-15: Swagelok Fittings............................................................................................. 73

Figure 4-16: Port Connector...................................................................................................75

Figure 4-17: Evacuation and Charging System Assembly.................................................. 76

Figure 4-18: Final LHP Assembly........................................................................................ 80

Figure 5-1: Schematic of Test Platform in Heat Pipe Laboratory....................................... 82

Figure 5-2: Thermocouple Placement for both Configurations.......................................... 84

Figure 5-3: Evaporator-Heat Source Assembly................................................................... 86

Figure 5-4: Condenser-Heat Exchanger Assembly 1...........................................................88

Figure 5-5: Condenser-Heat Exchanger Assembly 2 ...........................................................88

Figure 5-6: Addition of Subcooler to Liquid Line............................................................... 90

Figure 5-7: Baseline Results for 5 and 10W Power Inputs................................................. 93

Figure 5-8: Steady-State Evaporator Temperatures for a Given Power Input....................94

Figure 5-9: Temperature Profile for a 5 W Start-Up............................................................95

Figure 5-10: Power Cycling (5-10-15-10 W)....................................................................... 97

Figure 5-11: Power Cycling (2-20-2-20 W)......................................................................... 97

Figure 5-12: LHP Failure, 20 W Initial Input Power...........................................................98

Figure 5-13: LHP Failure, 10 W Initial Input Power......................................................... 100

Figure 5-14: 20 W Start-Up, 24 g Fluid Inventory............................................................ 103

Figure 5-15: Temperature Profile for 20 W Start-Up, 24 g Fluid Inventory................. 104

Figure 5-16: Temperature Profile for 20 W Start-Up, 15 g Fluid Inventory................. 105

Figure 5-17: 5 W Start-Up, 15 g Fluid Inventory.............................................................. 106



Figure 5-20: Temperature Profile for 20 g Fluid Inventory.............................................. 112

xii


Figure 5-21: Temperature Profile for 23 g Fluid Inventory...............................................113

Figure 5-22: Performance Curve for each Fluid Inventory................................................114

Figure 5-23: Effective Thermal Resistance for each Fluid Inventory...............................117

Figure 5-24: Heat Transfer Coefficient for each Fluid Inventory..................................... 119

Figure 5-25: Effect of Elevation on Steady-State Operating Temperatures.................... 121

Figure 5-26: LHP Operation During Periodic Heating. 15 W, 10 min cycle.................. 123

Figure 5-27: Temperature Hysteresis.................................................................................. 124

Figure 6-1: LHP Design Software Toolbar.........................................................................127

Figure 6-2: Edit Properties................................................................................................... 128

Figure 6-3: Working Fluid Properties Calculator.............................................................. 129

Figure 6-4: Figure of M erit.................................................................................................. 131

Figure 6-5: Fluid Inventory and Compensation Chamber Sizing..................................... 132

xiii


Nomenclature

Cp Specific heat, J/Kg-K

D Diameter, m

/ Darcy friction factor

g Gravity, 9.81 m/s2

h Height between the evaporator and condenser, m

k Thermal conductivity, W/m-K

keff Effective thermal conductivity of the wick, W/m-K

K mid Coefficient of free volume in the reservoir

L Length, m

•

m Mass flow rate, kg/s

M Mass of fluid to be charged, kg

AP Capillary pressure developed in the wick, N/m

APcond Pressure drop across the condenser, N/m

APf Frictional pressure drop, N/m2

AP Pressure drop/gain due to gravitational forces, N/m

^grooves Pressure drop in the grooves, N/m2

xiv

2


APu2

Pressure drop across the liquid line, N/m

Pressure drop across the sub-cooler, N/m2

APtot Total pressure drop in the system, N/m

APvl Pressure drop across the vapour line, N/m2

APwick Pressure drop in the wick, N/m

Qapp Total heat load applied to the evaporator, W

Q c c a Heat exchange between compensation chamber and ambient, W

Q e Net heat load applied to the evaporator, W

Qe,cc Heat leak from the evaporator to the compensation chamber, W

Qe,vap Heat used during the vaporization process, W

Q ll,a Parasitic heating of the liquid line from the surroundings, W

Q sc Amount of subcooling brought back by the liquid in liquid line, W

Qwick Heat flux through a porous wick, W

Veff Effective radius of the meniscus in the wick, m

R Local radius of the meniscus in the wick, m

R eff Effective thermal resistance of the LHP, °C /W

s Surface area, m2

Ta Ambient temperature, °C

Te Evaporator temperature, °C


W/K

Fluid temperature, °C

Compensation chamber temperature, °C

Saturation temperature, °C

Temperature of the liquid leaving the sub-cooler, °C

sink Sink temperature, °C

wall Wall temperature, °C

ATwick Temperature difference across the wick, °C

ydP^atSaturation temperature-pressure slope, °C /Pa

( UAl Thermal conductance between the evaporator and compensation chamber,

fUA^ Compensation chamber to ambient conductance, W/m-K

L )uFluid to ambient conductance, W/m-K

(U A '

\ J f-sink

Fluid to the sink conductance, W/m-K

Velocity, m/s

Volume of the compensation chamber, m

cond Volume of the condenser, m

grooves Volume of the grooves, m

xvi


V, Liquid volume, m3

Vlhp Total volume of the system, m3

Vn Volume of the liquid line, m3

vloop Total volume of the loop (excluding the compensation chamber), m"

Vp w

-2

Volume of the primary wick, m

K»■2

Volume of the secondary wick, m

Ki Volume of the vapour line, m3

a Void fraction of the compensation chamber

Hcond Heat transfer coefficient of the condenser, W/m -K

a e Heat transfer coefficient of the evaporator, W/m -K

P Fraction of compensation chamber volume occupied by the liquid

£ Wick porosity

^wick Wick permeability, m2

A Latent heat of vaporization, J/kg

M2 2Fluid viscosity, m /s

Av-2

Difference in the vapour and liquid specific volumes, m /kg

0 Contact angle, rad

P-2

Density, kg/m

<j Surface tension, kg/s

Nu Nusselt number

xvii


Pr

Re

Prandtl number

Reynolds Number

Subscripts

a Ambient

cond Condenser

e, evap Evaporator

eff Effective

f Fluid

i Internal

1 Liquid

11 Liquid line

0 External

sc Sub-cooler

V Vapour

vl Vapour line

xviii


Chapter 1: Introduction

A loop heat pipe (LHP) is a robust two-phase heat transfer device that uses the latent heat

of vaporization of a working fluid to transport waste heat between a source and a sink.

Similar to conventional heat pipes, LHPs are highly efficient devices that use the surface

tension developed in a porous wick structure to create the forces necessary for circulation

of the working fluid. LHPs also overcome some of the limitations imposed by

conventional heat pipes such as operability against gravity, use of smooth-wall flexible

transport lines, and fast diode action. Furthermore, LHPs are passive devices that require

no external power or moving parts for operation and are therefore highly reliable and

stable with a long operational life span. The presence of a secondary wick also provides

turnkey start-up and resistance to deprime.

Traditional LHPs consist of five main components: an evaporator, a compensation

chamber (or reservoir), a condenser, a vapour line, and a liquid line. Typically, only the

evaporator and compensation chamber contain porous wick structures while the rest of

the loop is made of smooth wall transport lines. A schematic diagram of a traditional

LHP is shown in Figure 1-1.

1


CHAPTER 1: INTRODUCTION 2

1-Vapour Grooves EvaporatorPrimary WickReservoir, Wick

End Cap

SecondaryWick

Evaporator Core

Liquid Line

Vapour Line

Condenser

Figure 1-1: LHP Configuration

The primary wick, located in the evaporator, is made of small pores for the purpose of

developing the necessary capillary pressure for circulating the fluid. In many LHPs a

secondary wick is employed between the evaporator and compensation chamber,

thermally and hydrodynamically connecting the two. The secondary wick is typically

made of larger pores to reduce the pressure drop of liquid flow and ensure liquid is

available to the primary wick at all times. The secondary wick also increases the

tolerance to bubble formation in the evaporator core, thus ensuring efficient loop

operation.

LHP operation is based on the same physical principles as those used in conventional

heat pipes, however, they are organized more efficiently (Maidanik, 2005). The



operation of a typical LHP begins with the application of heat to the evaporation zone.

Initially, the temperature of the evaporator and compensation chamber rise together.

Then, due to the thermal resistance of the wick, the temperature and pressure of the

evaporator begins to rise above that of the compensation chamber. LHP start-up occurs

once the temperature difference between the evaporator and compensation chamber is

high enough to initiate circulation of the working fluid. Start-up is easily identified by a

sudden increase in the temperature of the vapour line as superheated fluid flows through

the vapour line from the evaporator outlet to the condenser inlet. Once in the condenser,

heat is rejected and the vapour begins to condense into liquid phase. The condensed

liquid is then subcooled in the liquid line and returns to the evaporator through the

compensation chamber. The two-phase compensation chamber stores any excess fluid

and is responsible for establishing the loop operation pressure and temperature. No

external pumping power or moving parts are required for operation. Also, due to their

inherent flexibility, LHPs have the ability to be controllable, reversible, ramified,

flexible, and miniature, for a wide range of space and terrestrial applications.

1.1 Background

1.1.1 Conventional Heat Pipes

In 1944, Gaugler was the first to introduce the concept of a passive two-phase heat

transfer device with a wieking material. The idea re-surfaced in 1962 in connection with

the space program (Trefethen, 1962) and again in 1963 through a patent filed by Wyatt.



In 1964, the concept became more widely accepted as a means for thermal control after

Grover published results from an independent study. The publication by Grover was the

first to apply the term ‘heat pipe’ and described the apparatus as a ‘synergistic

engineering structure which is equivalent to a material having a thermal conductivity

greatly exceeding that of any known metal’ (Peterson, 1994). Since that time, heat pipes

have been employed in numerous applications ranging from temperature control of the

permafrost layer under the Alaskan pipeline to the thermal control of optical surfaces in

spacecraft (Peterson, 1994).

EVAPORATOR CONDENSERAdiabatic Region

1 1 1 1 1,1

7 WICK

LIQUID \VAPOR

Figure 1-2: Conventional Heat Pipe

A heat pipe consists of a sealed container lined with a wicking structure. The wick is

saturated with the liquid phase of a working fluid while the remaining volume (inner

core) contains the vapour phase. As illustrated in Figure 1-2, a heat pipe consists of three

distinct regions: an evaporator, a condenser, and an adiabatic region. Heat applied to the

evaporator by an external source vaporizes the working fluid in this section. The



resulting high pressure in the region drives the vapour from the evaporator to the

condenser where the vapour condenses and gives up its latent heat of vaporization to a

heat sink in this section. Capillary pressure, developed in the wick, then pumps the

condensed liquid back to the evaporator for the cycle to continue. This process continues

for as long as the flow of working fluid is not blocked and a sufficient capillary pressure

is consistently maintained (Chi, 1976).

1.1.2 Capillary Pumped Loops (CPL)

The capillary pumped loop (CPL) is very closely related to the LHP. The CPL was

invented in the United States in the 1960’s, but active development did not begin until the

early 1980’s when the National Aeronautics and Space Administration (NASA) Goddard

Space Plight Center sponsored Swales Thermal Systems and Dynatherm for a large

portion of the development and test effort (Nikitkin and Cullimore, 1998). The main

difference between a CPL and LHP is the location of the reservoir. In a traditional CPL,

the reservoir is usually connected to the liquid line by a reservoir line to store excess

liquid in the system and is sometimes thermally connected for temperature control

purposes. As a result, the fluid returning to the evaporator from the condenser does not

pass directly through the reservoir. In addition, the primary wick in a CPL is typically

made of polyethylene, with a low thermal conductivity and relatively large pore size.

These differences result in some operational differences between CPLs and LHPs. One

disadvantage of a CPL is that it requires preconditioning, where the reservoir is heated to

a temperature higher than the evaporator prior to start-up. In comparison, a LHP starts up



as soon as the temperature gradient between the evaporator and compensation chamber is

large enough to initiate fluid circulation. Another disadvantage of the CPL is its

tendency to deprime without sufficient subcooling. Without sufficient subcooling, the

evaporator core becomes blocked with vapour causing the wick to dry-out and the loop to

deprime. In the case of a LHP, a decrease in liquid subcooling is balanced by an increase

in loop operation temperature. Thus, LHP is ‘autoregulating’, ensuring stable and

continuous operation (Nikitkin and Cullimore, 1998). One major advantage of the CPL

in comparison to LHPs is its reservoir is relatively small and is separate from the

evaporator. This characteristic makes the unit much less sensitive to heat leak and much

more flexible for integration purposes. A detailed comparison between CPL and LHP

technologies is available in (Nikitkin and Cullimore, 1998) and (Ku, 1993).

1.1.3 Loop Heat Pipes (LHP)

The development LHPs originated from a pair of scientists from the former Soviet Union.

In 1972, at the Ural Polytechnical Institue of Thermal Physics, Gerasimov and Maidanik

successfully constructed and tested a device with a length of 1.2 m and capacity of

approximately 1 kW (Maidanik, 2005). In 1985, LHP design features and operation

principles were disclosed through patents filed by Maidanik under the name ‘Heat

Transfer Apparatus’ (Maidanik et al., 1985). It was in 1989 when the term ‘loop heat

pipe’ was first used, the year in which the first LHP flight test was conducted aboard the

Russian spacecraft Granat. During this flight test, long-term and reliable flight operation

in microgravity was successfully demonstrated (Kaya and Hoang, 1999). Further flight



testing experiments were conducted during the 1990’s including the flight test of an

American LHP on the STS-83 and STS-94 missions in 1997 (Lashley et al., 1998).

During these flights, the LHP demonstrated perfect on-orbit operation. The LHP has

since been widely accepted as a baseline technology for spacecraft thermal control

including that of the Boeing-702 bus used for Telesat’s ANIK FI and ANIK F2 satellites

as well as used for missions including NASA’s GLAS, EOS-Chemistry and GOES

spacecraft, ESA’s ATLID, CNES’s STENTOR, and RKA’s OBZOR (Ku, 1999).

LHP technology may enable further innovation in spacecraft thermal control such as

the use of deployable radiators or thermal control for human missions to Mars (Kaya and

Ku, 2003). LHPs may also offer excellent thermal control solutions for multiple

payloads onboard small satellites. Introduction of LHPs into other industries has been

slow because of their high fabrication costs. However, due to the recent research into

low cost wicks and manufacturing techniques, it is becoming more feasible to implement

LHPs into many advanced products for terrestrial application. LHPs have great potential

for use in a wide range of areas such as high power density chips, laptop computers,

avionics cooling, airplane anti-icing systems, nuclear power plants, road and bridge de-

icing, roof top solar installations, and remote communication sheds.

1.1.4 Miniature Loop Heat Pipe (MLHP)

Increased demand for components capable of handling higher heat load densities while

decreasing component mass and size has made it necessary to look for new approaches to

expand the potential application of LHPs. Therefore, more efficient designs with even



smaller dimensions are required. Recent efforts directed towards miniature loop heat

pipes (MLHPs) have yielded promising results. For example, Maidanik (2005)

investigated several MLHPs with masses ranging from 10-20 g and evaporator diameters

of 5-6 mm. Each MLHP was constructed with either stainless steel or aluminum and

used ammonia as the working fluid. In almost all cases, the units were capable of

transferring heat loads of up to 95 W over distances of up to 300 mm with a thermal

resistance ranging from 0.1 to 0.2 K/W. Of importance was that each unit possessed

sufficient mechanical flexibility, allowing for easy integration using a simple finned sink

and fan at the condenser. Khrustalev et al. (2001) investigated an ammonia/stainless steel

MLHP with a 5.6 mm outer diameter evaporator. The MLHP demonstrated successful

operation between a heat load range of 1 to 100 W with an overall thermal conductance

of about 5 W/K. An advanced feature in the design was that the temperature of the loop

was controlled using a thermal electric cooler (TEC) attached to the compensation

chamber. The temperature of the evaporator was maintained by either cooling or heating

the compensation chamber as necessary. Pastukhov et al. (1999) investigated two

ammonia/stainless steel MLHPs, both with a 6 mm outer diameter evaporator. Tests

were performed under atmospheric conditions with heat removal at the sink using both

free convection and forced convection. The first specimen possessed a heat transport

capacity of 50 W, but demonstrated unreliable start-up and operation capacity in

unfavourable orientations. The second MLHP tested showed capacity to provide reliable

start-up and stable operation with a heat transport capacity of up to 20 W in all



orientations. Khrustalev (2001) investigated the use of low cost MLHPs for use in

transporting, spreading, and dissipating heat in electronics. In particular, Khrustalev

looked at the use of a single copper wick sintered directly in the evaporator envelope.

The ‘hybrid’ wick structure was configured such that the liquid evaporates from the

capillary structure similar to that in a conventional heat pipe with a non-inverted

meniscus. The water/copper MLHP had an evaporator diameter of 9 mm and condenser

length of 100 mm. The MLHP transported more than 150 W at a steady-state

temperature of 75 °C, approximately one order of magnitude better than water/copper

conventional heat pipes.

1.2 Research Objectives

The main objective of this study was to gain an overall understanding of LHP technology

and to design and develop several functional units for testing. The purpose of

manufacturing an LHP was to demonstrate the technology’s reliability and flexibility for

use in many thermal control applications, and in particular, electronics cooling. Results

from testing each unit were used as a stepping stone for the development of much smaller

and more efficient designs. The following activities were considered to achieve the

above mentioned objectives:

• Establish a set of design requirements

• Develop procedures for component selection, manufacturing, cleaning, assembly,

evacuation and charging

• Equip the LHP laboratory with capabilities for LHP fabrication and testing



• Establish an experimental approach and testing procedures

• Identify key areas of concerns to be addressed through performance testing

• Develop mathematical models to estimate working fluid properties, and calculate

fluid inventory and compensation chamber sizing

Tests were performed on each LHP to experimentally study critical performance

characteristics. The tests were carried out in three phases as follows:

1. Investigate LHP functionality

2. Investigate sensitivity of LHP performance with respect to fluid charge; measure

steady-state operating temperature, maximum transport capacity, and effective

thermal resistance.

3. Study the effect of elevation and periodic heating on the maximum transport

capacity and stability of the LHP. Investigate temperature hysteresis

characteristics.

1.3 Organization

This thesis presents a detailed discussion and analysis of the design and development of

loop heat pipes and is organized as follows:

Chapter 1 Introduction: Introduces the concept of LHPs and provides a brief history

of the technology and an overview of the main research objectives and organization

of the thesis.

Chapter 2 Literature Review: Presents LHP operation principles and characteristics

and discusses traditional design processes.



Chapter 3 LHP Design: Introduces the problem statement and design approach. Also

presents a detailed discussion of component sizing and selection.

Chapter 4 LHP Fabrication: Provides a detailed overview of the manufacturing

process including cleaning, assembly, evacuation, and charging procedures.

Chapter 5 LHP Testing: Presents the experimental approach, setup, and procedure.

Test results of each unit are also presented and examined.

Chapter 6 Mathematical Modeling: Discusses the development of mathematical

models used to estimate working fluid properties, fluid inventory, and compensation

chamber volume.

Chapter 7 Conclusions and Recommendations: Summarizes the work completed,

provides conclusions, and discusses future work and recommendations.


Chapter 2: Literature Review

2.1 LHP Operating Principles

The primary wick in the evaporator must develop a capillary pressure sufficient to

overcome the total pressure drop in the loop to ensure continuous LHP operation. The

total pressure drop in the system is the sum of the frictional pressure drops in the vapour

grooves, vapour line, condenser, liquid line and evaporator wick plus any static pressure

drop due to gravity (Ku, 1999):

APf . = AP + AP.+AP , + AP„ + A P ., + AP + AP (2.1)tot grooves vl cond 11 w ick sc g V /

For fully developed single-phase flow, the frictional pressure drop can be calculated with

the following equation:

A / WpV

v 2 ,(2 .2)

where / is a dimensionless parameter called the Darcy friction factor and is a function

of the Reynolds number and tube roughness. The pressure drop across the wick structure

is characterized by the liquid viscosity and wick permeability and may be approximated

as:

12


CHAPTER 2: LITERATURE REVIEW 13

wick11 m

2^ KWickPlIn (2.3)

where Kwick is the permeability of the wick, ju, is the liquid viscosity, and m is the mass

flow rate of the working fluid. The gravity pressure losses are the result of hydrostatic

pressure and are not temperature dependent. These losses are defined as:

APg =pgh (2.4)

One of the main advantages of any capillary loop is that the meniscus inside the wick

automatically adjusts itself by reducing the radius of curvature, thus establishing a

capillary pressure equal to that of the total system pressure drop. The capillary pressure

developed by the wick is expressed by the Young-Lap lace equation:

AP = 2ac0s(> (2.5)cap R

where R is the radius of curvature of the meniscus in the wick, a is the surface tension

of the working fluid, and 6 is the contact angle between the liquid and solid. As the heat

load is increased, the radius of curvature will continue to decrease until it is equal to the

effective pore radius of the wick, reff. At this point, the wick has reached its maximum

capillary pressure. The maximum pressure that a wick structure can develop depends on

design and size and is defined as:

AP = 2aCOs(> (2.6)cap,m ax VArf*vvrreff



If the total system pressure drop exceeds the capillary limit that the wick can provide,

then the wick will experience ‘dry-out’. In this situation vapour penetrates the evaporator

core through the wick, and a sudden and sharp increase in operating temperature is

observed. Fortunately, as evident in many designs, LHPs have the ability to recover from

dry-out simply by lowering the heat load (Ku, 1999).

2.1.1 Thermodynamic Cycle

The steady-state operation of LHPs may be better understood through a simple

thermodynamic analysis of a pressure-temperature diagram. Figure 2-1 illustrates the

thermodynamic cycle upon which a typical LHP operates. The thermodynamic states

indicated by the numbers in Figure 2-1 also correspond to the physical LHP locations

shown in Figure 1-1. In the analysis to follow, it is assumed that the liquid and vapour

lines are perfectly insulated. In reality however, there is usually some heat exchange with

ambient through either convection or radiation. Correspondingly, the temperature in each

line may increase, decrease, or stay the same depending on the heat exchange and

temperature of ambient.



<DDWCO0L _

LIQUID

CL

VAPOUR

Temperature

Figure 2-1: Pressure-Temperature Curve of Steady-State LHP Operation

As illustrated in Figure 2-1, point 1 represents the conditions at the evaporator surface

(on the vapour side of the meniscus). As thermal energy is added to the working fluid,

vapour is generated at the outer diameter of the evaporator wick and is at saturation

temperature corresponding to the highest pressure in the system. As the vapour flows

along the vapour grooves to the evaporator exit (point 2) it becomes slightly superheated

due to a small decrease in absolute pressure and an increase in temperature caused by

contact with the evaporator wall. As the vapour flows along the vapour line, between

points 2 and 3, temperature remains constant while the pressure continually decreases due

to frictional losses. At the end of the vapour line (point 3) the vapour becomes more



superheated relative to the local saturation pressure. While in the condenser, heat from

the vapour is rejected to the sink and the vapour begins to condense. Condensation of the

vapour continues, with both pressure and temperature decreasing, until condensation is

complete and the fluid reaches a saturated liquid state on the saturation line (point 4)

which is often referred to as the liquid-vapour interface in the condenser. Assuming the

condenser is not fully utilized, the liquid begins subcooling inside the condenser until it

reaches the liquid line inlet at point 5. In the liquid line, the temperature remains constant

while the pressure gradually decreases along the line (points 5 to 6). Points 6 to 7

represents the liquid motion in the compensation chamber and evaporator core. Since the

evaporator core is considered an extension of the compensation chamber, both have the

same absolute pressure. Also, since the evaporator core is thermally and hydraulically

connected to the compensation chamber, the temperature and pressure at the inner surface

of the wick should be at saturation (point 7). Points 7 to 8 correspond to the liquid

motion through the wick structure and back to the evaporation zone. Capillary action

draws the liquid at the inner surface of the wick to the outer surface of the wick. As the

fluid travels through the porous wicking material, it gains in temperature while also

decreasing in pressure. Point 8 determines the state of the working fluid in the vicinity of

the evaporating menisci and has the lowest pressure in the system (Chuang, 2003).

The difference in pressure between the evaporator and compensation chamber

provides the driving force for the circulation of the fluid. Since the two saturation states



are thermodynamically related, the following condition must be satisfied for steady-state

LHP operation:

/P - P =tot wick ydT j

More specifically, Eq. (2.7) can be written as:

ftp( T . - T j ) (2-7)

P - P =i Cf( d P \

(2-8)

where:

, Tj is the saturation pressure and temperature in the vapour grooves

Rj, T7 is the saturation pressure and temperature in the compensation chamber

dP— is the slope of the saturation line at the corresponding pressure anddT

temperature of the compensation chamber

Furthermore, the derivative of dP/dT is related to the physical properties of the working

fluid by using the Clausius-Clapeyron equation:

dP XdT TAv

where:

X is the latent heat of vaporization of the working fluid

T-j is the saturation temperature of the fluid in the compensation chamber

Av is the difference in the vapour and liquid specific volumes

(2.9)



Eq. (2.8) illustrates that for a given pressure gradient between the evaporator and

compensation chamber, there must be a corresponding difference in saturation

temperatures. Furthermore, if external conditions change, the loop operating temperature

will move along the saturation line until a new equilibrium is reached. Ku (1999)

suggested that the demonstrated coupling between the pressure and temperature across

the wick has implications on loop operating temperature at low heat loads and at adverse

orientations.

2.2 I.HP Operating Characteristics

2.2.1 Loop Operating Temperature

As previously discussed, the compensation chamber is located adjacent to the evaporator

and is thermally and hydraulically connected by the primary and secondary wick. Also,

the working fluid must pass through the compensation chamber from the condenser to the

evaporator. For this reason, the temperature of the compensation chamber is a function

of evaporator heat load and enthalpy of the subcooled fluid returning from the condenser.

The condenser is a function of the condenser sink temperature and ambient temperature.

As will be shown, the compensation chamber is the primary driver for the loop operating

temperature. A simplified thermal network of the LHP model is shown in Figure 2-2 (a

complete thermal network can be found in (TAIS, 2000)). The model is used to better

understand the complex process.



Ta

Q 1cc,a I

m,hm,h m.he

vl ,a

■—wwr*

m,h m,h,? vl

T~Tsat

Liquid Line Vapour Line

'sc,s ink cond,sink

Sink Temperature

m,h .* cond L

cond "^"sat

AdiveCondenser

In adive Condenser

CompensationChamber Evaporator

Figure 2-2: Simplified Thermal Network of the LHP Model

As shown in Figure 2-2, the compensation chamber exchanges energy with the

evaporator, ambient, and subcooled liquid returning from the condenser. As heat is

applied to the evaporator, part of it goes directly towards vaporization of the working

fluid while the other part Teaks’ to the compensation chamber and is expressed as

follows:

Qe = Qe,cc + Qe.yap (2-10)

where Qe is the net heat load applied to the evaporator. The term, Qe vap, is the heat used

during the vaporization process of the working fluid and can be expressed as:



Any heat transfer from the evaporator to the compensation chamber is referred to as heat

leak and is expressed as:

Q . ,« = ( U A l j T , - T j (2.12)

where {UA) is the thermal conductance between the evaporator and the compensation

chamber. The amount of heat leak is proportional to the heat load and depends strongly

on the flow conditions inside the evaporator (Chuang, 2003). More specifically, when

the evaporator core is completely filled with liquid, heat transfer between the

compensation chamber and evaporator is primarily through axial conduction of the

evaporator body. However, if vapour is generated at the inner surface of the primary

wick, it may be transported back to the compensation chamber thus increasing the heat

leak through a process similar to a conventional heat pipe (ie. evaporation and

condensation of the bubbles between the evaporator core and compensation chamber).

At steady-state, the heat leak from the evaporator must be balanced with the heat loss

to ambient and the amount of subcooling that is brought back by the liquid from the

condenser. Since the heat exchange between the compensation chamber and ambient is

by natural convection, it is relatively small compared to the amount of subcooling

brought back from the liquid line. Therefore, by assuming the heat exchange between the

compensation chamber and ambient is negligible, the following must apply:

Q„c =mCr&.T = mCr (Tcc-T„) (2.13)



where Tin is the liquid temperature at the entrance to the compensation chamber and AT

is the amount of liquid subcooling. As the liquid exits the condenser and flows through

the liquid line, it will exchange heat with its surroundings. The temperature difference

caused by the heat exchange with ambient is expressed as:

T » - T „ = ^ (2.14)mCp

where Tsc is the temperature of the liquid leaving the subcooler and Qll a is the parasitic

heating of the liquid line from the surroundings. At relatively low heat loads, a very

small mass flow rate is produced and only part of the condenser is utilized for

condensation. Consequently, as the liquid flows along the liquid line, its temperature is

raised close to that of ambient due to parasitic heating. Decreased subcooling in the

liquid line must be compensated for by an increase in the compensation chamber

temperature in order to balance the heat leak from the evaporator. Demonstrated through

Eq.(2.14), it is observed that a very large temperature increase may be required due to a

very low mass flow rate. As the heat load is increased, the mass flow rate also increases

and the liquid spends less time in the liquid line thus minimizing parasitic heating. By

increasing both the mass flow rate and the liquid subcooling, the temperature of the

compensation chamber decreases. This process continues until the condenser is fully

utilized and only vapour exists. At this point, the compensation chamber temperature is

at a minimum.



This process is demonstrated using the performance curve of a classical LHP, as

shown in Figure 2-3 for a sink temperature below ambient temperature. The figure is

separated into two parts: variable conductance mode and constant conductance mode. In

variable conductance mode, at very low powers, only a small section of the condenser is

active while the rest is used to cool the liquid phase of the working fluid. As the power is

increased, more condenser area becomes active thus increasing the conductance of the

condenser and the overall conductance of the LHP. At a certain power, Qt, the condenser

is completely active and further increases in the overall conductance are no longer

possible. Therefore, when the heat load is greater than or equal to Qt, the LHP operates

in fixed conductance mode. At this point, the vapour-liquid interface fluctuates near the

end of the condenser line. Under this condition, since the condenser conductance and

sink temperature are constant, the steady-state operating temperature increases linearly

with increasing heat load (Nikitkin and Cullimore, 1998). The transition from variable to

fixed conductance modes is dependent on sink temperature and the thermal coupling

between the LHP and the environment (MacDonald, 2004).



FixedConductance Mode /

V ariableConductanceMode

Oo

0Q.E0H

Applied Heat Load

Figure 2-3: Variable and Fixed Conductance Modes

2.2.2 Start-Up

One of the primary advantages of an LHP is its reliable turnkey start-up. Unlike CPLs,

LHPs do not require preconditioning. As heat is applied to the evaporator of an LHP, the

working fluid circulates within the loop thus removing heat from the evaporator to the

condenser. However, there exists a minimum power load in which an LHP will ‘start

up’. The heat load must be sufficient to produce the required pressure difference (due to

a temperature gradient between the compensation chamber and evaporator) to initiate

circulation of the fluid. It has been found that many LHP devices experience start-up

problems at very low heat loads (usually 10W or less) or in the presence of heavy masses

attached to the evaporator. Nikitkin et al. (1998) suggested that under these conditions, a



Peltier element might be needed to slightly and temporarily chill the compensation

chamber to produce the required temperature gradient to start circulation. Similarly, an

LHP can be stopped by heating the compensation chamber above the temperature of the

evaporator. The minimum heat load for start-up is dependent on a number of factors

which include the evaporator and compensation chamber design, initial conditions in the

evaporator, and operating conditions prior to start-up. Most importantly, the initial state

of the working fluid in the evaporator and across the wick greatly impacts LHP start-up

characteristics including temperature overshoot and minimum power requirements.

Ku (1999) demonstrated four typical configurations of the liquid and vapour states

inside the evaporator and compensation chamber prior to start-up (see Figure 2-4). He

proposed that there are three major factors affecting start-up. First, if liquid completely

fills the vapour grooves, a liquid superheat will be required to initiate boiling. However,

if the vapour grooves already contain vapour, then boiling will be initiated as soon as

power is applied and without any liquid superheat. Second, if the evaporator is

completely liquid-filled, heat exchange between the evaporator and compensation

chamber due to heat leak is minimized since heat is transferred primarily through

conduction. However, if there is vapour present in the evaporator core, heat leak is

significantly increased as the evaporator core becomes a vapour extension of the

compensation chamber. Finally, the applied power affects start-up through interactions

with the other two factors.



i f I I M i f

(a) Vapour channel: two-phase

Evaporator core: liquid-filled

(b) Vapour channel: liquid-filled

Evaporator core: liquid-filled

(c) Vapour channel: two-phase

Evaporator core: two-phase

.(d) Vapour channel: liquid-filled

Evaporator core: two-phase

Figure 2-4: LHP Start-up Scenarios in the Evaporator

Experimental analysis of LHP start-up under varying conditions can be found in

literature. For example, Cheung et al. (1998) presented experimental data suggesting that

superheat of the liquid is significantly reduced when two-phase fluid exists in the vapour

channels. Similarly, Kaya et al. (1999) demonstrated start-up characteristics at different

orientations. They showed that the required superheat, maximum temperature at start-up,

and time required for start-up strongly depend on loop orientation and therefore the

overall pressure drop. They also proposed that the presence of vapour bubbles in the

vapour grooves at high elevations promote nucleate boiling and as a result reduce the

required superheat and time for start-up.



2.2.3 Temperature Hysteresis

Temperature hysteresis is a phenomenon that occurs when the steady state operating

temperature of an LHP depends not only on the applied heat load but also on the recent

history of the heat load sequence for the same operating conditions such as sink

temperature, ambient temperature and orientation. The typical trend of temperature

hysteresis is shown in Figure 2-5 (Chuang, 2003) below.

Oo**wCDi _32CDCLE0I—O)|Z+->CS1—0CLo

Temperatu re Hysteresi s

Normal Operation

Q t

Applied Heat Load (W)

Figure 2-5: Typical Temperature Hysteresis Trend

As indicated in Figure 2-5, temperature hysteresis is dominant in the low power range

and is associated with large decreases in applied power. This phenomenom may be

explained as follows. As the heat load to the evaporator decreases, liquid from the

compensation chamber is injected in the condenser. This process shifts the vapour-liquid

interface towards the condenser entrance, reducing the area of condensation. For a



moderate power decrease, the secondary wick ensures only liquid is supplied. However,

with a large power decrease the pressure required to transfer liquid may exceed the

capillary limit of the secondary wick. In this case, vapour bubbles may accumulate in the

evaporator, changing the void fraction of the evaporator core. As the void fraction in the

core increases, so to does the heat leak to the compensation chamber. Therefore, as more

and more vapour is introduced into the evaporator core, the operating temperature of the

loop increases since the subcooling term (m CpAT) can only compensate for the

increased heat leak by increasing A T . In the high power range, above Qt, a higher mass

flow rate increases liquid subcooling thus collapsing excess vapour bubbles inside the

evaporator core and eliminating temperature hysteresis (Ku, 1999).

According to Kaya and Ku (1999), the temperature difference between the higher and

lower trend strongly depends on the individual LHP design and in some cases may not

even be present. Since the true mechanism of temperature hysteresis is not well

understood, LHPs should be thoroughly tested for temperature hysteresis before

implementation.

2.2.4 Effect of Orientation

As mentioned in section 2.2.1, heat leak from the evaporator to the compensation

chamber greatly affects the operating temperature of a LHP. Therefore, LHP orientation

in 1-g environments can significantly influence loop temperature by promoting heat leak

due to vapour bubble accumulation in the evaporator due to buoyancy forces.



Using mathematical models, Chuang (2003) established that orientation has little

effect on loop temperature at high heat loads because frictional pressure drops, which are

the result of a high mass flow rate, dominate hydrostatic losses. In contrast, low heat

loads produce lower mass flow rates, thus, gravitational head becomes a more dominant

factor. These results were experimentally demonstrated by Wolf et al. (1994) using an

ammonia LHP at two adverse elevations of 0.91 m and 2.74 m. In their experiment, they

found that LHP operating temperature increases with increasing adverse elevations for

low to moderate power inputs. Similarly, Ku (1999) explains that the increase in LHP

operating temperature due to an increase in adverse elevation is the result of an increase

in pressure difference across the wick. An increase in pressure difference across the wick

must result in an increase in saturation temperature difference across the wick (as

required by Eq.(2.7)). Since the enthalpy of the liquid entering the compensation

chamber does not change, the evaporator vapour temperature must increase to satisfy the

increasing pressure drop. This is turn leads to an increase in heat leak between the

evaporator and compensation chamber. The compensation chamber temperature must

then increase in order to provide enough liquid subcooling which in turn results in an

even higher evaporator vapour temperature. As a result, a dramatic increase in operating

temperature may accumulate rapidly. However, since hydrostatic pressure losses are

independent of flow rate, the effect of an increased adverse elevation may be suppressed

by increasing the evaporator heat load. Increasing the evaporator heat load increases the

mass flow rate in the loop which increases the liquid subcooling, sufficient to balance



heat leak. Therefore, at high heat loads, any increase in operating temperature due to

adverse orientation becomes unnoticeable.

2.2.5 Effect of Non-Condensable Gases

Significant non-condensable gases (NCGs) can sometimes be generated by improper

cleaning procedures before assembly, impurity of the working fluid, and chemical

reactions between the working fluid and case materials. The amount of NCGs generated

is a function of the amount of working fluid, surface area of the materials in contact,

operating conditions, and period of exposure (Ku, 1999).

The effect of NCGs on performance depends on the location of accumulation of the

gases. The NCG typically collects in the compensation chamber or condenser. It may

also circulate with the working fluid or be absorbed in the wick. In large quantities

NCGs can sometimes increase the start-up time and operating temperature of the loop.

This is accomplished if the NCG blocks part of the condenser thus reducing the overall

thermal conductance. Presence of the gases in the evaporator core may also increase heat

leak. However, Nikitkin and Bienert (1998) concluded that LHPs are relatively

insensitive to the presence of NCGs after test results indicated that the effect on LHP

performance is usually minimal.

2.3 Design of Loop Heat Pipes

Loop heat pipe design is an extremely complicated process involving a wide range of

variables such as size, mass, shape, volume, working fluid, wick material, and case



material. LHP design also involves such aspects as thermal load, transport length,

evaporator/condenser length, operating temperature range, source-sink interfaces, fluid

inventory, and life/reliability (Peterson, 1994). The design process can be streamlined by

formulating an iterative procedure as outlined in Figure 2-6.

ProblemSpecification

WickProperties

Fluid and Material

Properties

OPTIMUMSOLUTION

EvaluationCriteria

►

MathematicalModeling

OptionalSolutions

ExperimentalValidation

Preliminary Selection of Working Fluids Wick Materials Case Materials

Figure 2-6: Loop Heat Pipe Design Procedure Adapted from (Chi, 1976)

The first step is to clearly identify the problem statement and requirements. This may

include defining constraints for physical properties such as mass and external volume. It



may also include specifying thermal performance requirements such as maximum

transport capacity or operating temperature range. For example, it may be necessary to

design a LHP for a limited volume application such as laptop cooling where operational

temperatures are limited to -20 to 60 °C. Other requirements may include maximum heat

load and heat flux, mode of heat rejection, and start-up at low power inputs. The second

step involves preliminary selection of a working fluid, wick material, and case material.

In this step, it is necessary to select components which are compatible with each other.

Several combinations may exist that satisfy the requirements specified in the problem

statement. Some combinations of case materials, wick materials, and working fluids

have already been successfully applied in loop heat pipes. Examples of these

combinations are listed in Table 2-1.

Table 2-1: Tested LHP Material-Working Fluid Combinations (Maidanik, 2003)

Case Material Wick Working FluidStainless Steel Nickel Water, Ammonia,

Acetone, Pentane, Freon-152 A, Freon

11, PropyleneStainless Steel Titanium Water, Ammonia,

Acetone, Pentane, Freon-152 A, Toluene

Stainless Steel Stainless Steel Ammonia

Nickel Titanium AmmoniaNickel Nickel Ammonia

Copper Copper Water

It is suggested to use similar combinations as a starting point. The effect of each

component on operational characteristics must be determined before final selection.



Table 2-2 provides some indication of how each of the three primary components affects

various design requirements (Peterson, 1994).

Table 2-2: Components and their Influence on Design Requirements (Peterson, 1994)

Design Requirements WorkingFluid

WickMaterial

CaseMaterial

Thermal PerformanceTransport Capacity S S W

Operating Temperature Range S w wTemperature Drop M w w

MechanicalPhysical Requirements (size, W w Mweight, etc.)

Wall Thickness - internal pressure W N SSink - source interface N N sDynamic / static loads W S M

Reliability and SafetyMaterial Compatibility S S S

External Corrosion N N SFabrication M M M

Pressure containment/leakage W M SToxicity S W w

Gravitational Environment

>ig s S sIg M M w<ig W M w

S = strong factor; M = moderate factor; W = weak factor; N = negligible

The third step involves the use of mathematical models, such as the one proposed by

Kaya and Hoang (1999), to predict primary heat transfer characteristics. More

specifically, steady-state energy balance and pressure-drop equations are modeled along

the flow path of the LHP and used to estimate maximum heat load, system pressure

losses, and operational temperatures for the selected components and properties in step



two. The results are then verified against the requirements specified in the problem

statement. Steps one through three are repeated several times to yield optional solutions

for the specified design requirements (ie. varying combinations of each LHP component

are evaluated). Finally, evaluation criteria such as cost and manufacturing techniques are

input into an evaluation procedure to determine the optimum solution. The evaluation

procedure includes experimental testing for performance verification of initial designs.

It should be noted that some mathematical models have been validated with a good

degree of correlation to experimental values. However, some uncertainties related to

their design still remain (Riehl and Dutra, 2005). For example, phenomena such as

temperature hysteresis are not yet completely understood. Also, several simplifying

assumption are typically required in the development of these models. Therefore, it may

be necessary to manufacture and test several prototypes before an optimum design is

obtained, provided all necessary requirements are met.

2.3.1 Working Fluids

Many factors affect the selection of an appropriate working fluid, including operational

temperature range, vapour pressure, thermal conductivity, stability, toxicity, and

compatibility with wick and case materials. Different types of working fluids have been

studied and are well documented in literature for their use in conventional heat pipes.

Traditionally, working fluids have been categorized as either: cryogenic fluids such as

helium, neon, oxygen, and nitrogen; moderate-temperature fluids such as methanol,

ammonia, acetone, and water; or high temperature-liquid metal fluids such as potassium,



lithium, or sodium (Chi, 1976). Table 2-3 illustrates the typical operating temperature

ranges for various working fluids.

Table 2-3: Operating Temperature Range of Various Fluids Adapted from (Faghri, 1995)

Working FluidMelting point, K at 1 atm

Boiling point, K at 1 atm

Operating Temperature

Range, K

Classifiedtemperatureapplication

Helium 1 4.2 2-4Hydrogen 13.8 20.4 14-31

Neon 24.4 27.1 27-37Nitrogen 63.1 77.4 70-103 Cryogenic

Argon 83.9 87.3 84-116Oxygen 54.7 90.2 73-119Krypton 115.8 119.7 116-160

Ammonia 195.5 239.9 213-373Pentane 143.1 309.2 253-393

Freon 113 236.5 320.8 263-373Moderate

Acetone 180 329.4 273-393 TemperatureWater 273.1 373.1 303-473

Methanol 175.5 338.2 273-513Ethanol 158.8 351.4 273-513

Sodium 371 1151 873-1473Lithium 453.7 1615 1273-2073 High Temperature

Silver 1234 2385 2073-2573

LHPs have been designed primarily for use in space and electronic applications. LHP

applications in these areas require the selection of a working fluid with boiling

temperatures between 250 and 375 K. This typically limits the selection to fluids such as

ammonia, acetone, methanol, water, and Freon-11 or 113.

For efficient loop operation, working fluids should possess the following

characteristics: high latent heat of vaporization for more efficient heat transport, high



thermal conductivity to minimize temperature drops across the wick, high surface tension

to maximize capillary pumping capabilities, and low viscosity to minimize pressure

losses along the fluid flow line. Chi (1976) combined these properties into what is

known as the merit number or liquid transport factor to evaluate the effectiveness of

various working fluids at specific operating temperatures. The merit number is

calculated as follows:

N i = Pi L (2.15)Mi

where:

Pi is the liquid density

a is the surface tension

k, is the liquid thermal conductivity

Mi is the liquid viscosity

As an example, a figure of merit for some moderate to high-temperature working fluids is

provided Figure 2-7.



/ \o\

\± . .

T5 ,1 M e*hono l

200 400 600 f;ro u~J:\ '"""'1200 " 1400femperatuiL, V'

Figure 2-7: Figure of Merit for Various Working Fluids (Dunn and Reay, 1982)

Another concern regarding the selection of a working fluid is the compatibility

between the fluid and case material. Any chemical reaction between the working fluid

and case material creates non-condensible gases in the loop. The existence of NCGs may

degrade LHP performance. Compatibility of various metals with working fluids can be

found in literature (Faghri, 1995).

2.3.2 Primary and Secondary Wicks

The selection of the primary wick structure is an important step in the design process of

LHPs. The primary wick is located in the evaporator section of a LHP and provides the

required pressure that circulates the fluid in the system. The most critical properties to

consider when selecting the wick structure are: effective pore radius, porosity,



permeability, and effective thermal conductivity. Effective pore radius affects the

capillary limit of the wick while porosity and permeability affects pressure drop due to

fluid flow across the wick. Ideally, wick structures are manufactured to have an effective

pore radius of 1 -1 0 /xm, a porosity of 50-75% , and a permeability of greater

thanlxlO“13m2. Effective thermal conductivity affects heat leak from the evaporator to

the reservoir and should be minimized. Unfortunately, obtaining all of these properties

simultaneously in one sample can be quite difficult since they are often contradictory.

Therefore, some compromise is required when designing the wick structure. Another

important aspect to consider is the overall cost. Traditionally, metal wicks such as

sintered nickel, titanium, or copper powders are used in LHPs and are often the most

expensive component. This is due to the fact that the wick has to be manufactured in an

energy-intensive process where it is sintered, machined, evaluated, cleaned, inserted into

evaporator, and sealed before the entire assembly of the LHP (Khrustalev, 2001). Other

properties to consider include compatibility with working fluids and case material,

material availability, and machineability.

The secondary wick is used to continuously provide fluid from the compensation

chamber to the reservoir. Since its primary function is to supply fluid and not provide

high capillary pressure, a much larger effective pore radius (2 0 - 150 /jm ) is acceptable.

Therefore, metal mesh structures are typically used.



2.3.3 Component Sizing

Sizing of the evaporator and wick structures strongly influence the maximum heat load

and pressure drop of the loop, while the length and size of the transportation lines directly

influence the heat transfer between the working fluid and ambient, pressure drop of the

system, and size of the reservoir. Traditionally, LHPs have employed cylindrical

evaporators ranging from 12 to 28 mm in diameter. The shape and size of the condenser

can vary significantly depending on the means and conditions of cooling. The length of

vapour and liquid lines can reach up to 10 m while their diameter, as a rule, is in the

range of 3 to 8 mm (Pastukhov et al., 2003). Demand for smaller devices operating in the

10 to 100 W ranges has pushed the development of MLHPs and has also led to a

significant reduction in component sizes. For example, the length of heat transfer is

typically 0.5 to 1 m. Furthermore, the evaporator length is usually no more than 40 mm

while the evaporator diameter is typically no more than 6 mm. Finally, vapour and liquid

lines are approximately 1 to 2 mm in diameter.

2.3.4 Fluid Inventory and Compensation Chamber

Sizing the compensation chamber and estimating the fluid charge are critical tasks in

designing an LHP. Both affect a wide range of performance characteristics that can

include overall heat conductance, steady-state operating temperature, maximum power,

and minimum start up power. Generally, there are three design conditions to consider

when developing an LHP: cold operation, hot operation, and maximum non-operating

temperature. During cold operation, no heat is applied to the evaporator while the



transport lines (condenser, liquid line, vapour line) are exposed to the coldest

environmental conditions. During hot operation, maximum power is applied to the

evaporator while the rest of the loop is exposed to the hottest environmental conditions.

The maximum non-operating temperature is important during storage and/or

transportation where the concern is that the LHP may burst due to hydrostatic pressure at

elevated temperatures when there is not enough void volume in the loop. Generally,

there must be enough liquid to supply the evaporator for start-up during cold case

conditions while also preventing condenser blockage during hot case conditions.

Furthermore, the compensation chamber should be sized to compensate for the thermal

expansion of the working fluid at the different operating temperatures. It is therefore

important to size the compensation chamber and fluid inventory concurrently.

Several approaches found in literature have been developed to effectively select a

fluid charge and compensation chamber volume that would yield optimum performance

characteristics. The techniques documented by Ku (1999) and TAIS Ltd (2000) are

summarized below.

In the first approach, documented by Ku (1999), there is no theoretical upper limit for

the compensation chamber volume. However, the volume should be optimized with the

fluid inventory to accommodate space and weight constraints. Furthermore, the

minimum compensation chamber volume must be able to accommodate at least the swing

volume between the hot and cold case of loop operation. During cold operation, some



small portion of the compensation chamber is liquid-filled while the rest of the loop is

completely flooded. The fluid inventory must therefore satisfy the relation:

M = P,f (V,„ +0-V«) + p „ ( l - f i ) V " (2.16)

where:

pl c is the cold-case liquid density

p v c is the cold-case vapour density

Vlogp is the loop total volume excluding the compensation chamber

P is the fraction of compensation chamber volume occupied by the liquid

During hot operation, some vapour space should be available in the compensation

chamber when the condenser is completely open. Thus, the fluid inventory must also

satisfy the relation:

M = Pu [v„ + v„ + v„ +■(i ■-.a ) ■r„ ] + p,t (vsnm, + r„ + r „ „ +a-Vcc) (2.17)

where:

p, h is the hot-case liquid density

p v h is the hot-case vapour density

a is the void fraction of the compensation chamber

Ku (1999) states that a and P values are selected at the designer’s discretion and that a

careful selection of these two values will yield an optimal compensation chamber volume

and fluid inventory. Finally, the fluid charge needs to be checked against the upper limit



of the loop, imposed by the maximum non-operating temperature condition. In this case,

the fluid inventory must satisfy the following relation:

where p l max is the liquid density at the maximum non-operating temperature.

The second technique, documented by TAIS Ltd (2000), is a more conservative

approach that assumes there is always sufficient liquid in the LHP for start-up. That is,

LHP start-up will always occur by supplying heat to the evaporator, even in the most

unfavourable conditions and orientations. Therefore in this case, the amount of liquid in

the internal volume shall be sufficient to fill the entire LHP except the reservoir and

vapour grooves, as illustrated in Figure 2-8.

(2.18)

Reservoir andWick filled with linuid vapour grooves

iquid

Vapour and liquid lines filled with liquid

Condenser filled with liquid

Figure 2-8: Fluid Distribution Scheme 1



At maximum operational temperature, the compensation chamber must be sized such that

it can compensate for the volume expansion of the fluid (from worst case cold to worst

case hot conditions) as well as a shift of liquid from the vapour line and entire condenser.

The following algorithm is used to concurrently obtain the compensation chamber

volume and mass of fluid to be charged.

1. Initial approximation of working fluid volume expansion due to temperature

difference between hot and cold cases:

AVr = 0 (2.19)

2. Estimate compensation chamber volume:

K = ( w , + r ^ + r « ) - K M (2.20)

3. Calculate mass of fluid to be charged for minimum operation temperature:

M = (Kc + grooves )' Pv + Vlhp~{Kc + grooves ) ' Pl (2-21)

4. Calculate the liquid phase of the working fluid at minimum operational

temperature for the mass calculated in step 3:

y = M ~ VihP ' P, q 2T)Pl Pv

5. Calculate the liquid phase of the working fluid at maximum operational

temperature for the mass calculated in step 3:

V, = M ~ V,hp 'P' (2.23)Pl ~Pv

6. Calculate new value for volume expansion:



A V, = V, -V, (2.24)1 h m ax h m in

7. Repeat steps 3 through 6 until the desired accuracy is reached

KVoid is the coefficient of free volume in the compensation chamber (ie. the portion of the

compensation chamber that remains liquid-free during operation). Typical values range

from 1 to 1.2, where 1.2 corresponds to 17% free volume in the compensation chamber.

Variants of this method can be used to minimize the amount of working fluid necessary

for operation. Such cases may exist if start-up is controlled through active cooling of the

compensation chamber or if the LHP is oriented in such a way that the evaporator already

has the necessary liquid supply. Additionally, it may be required that the condenser

remain completely open throughout the entire LHP operating range in applications where

constant conductance is required. It should be noted, however, that this may introduce

some degree of risk with an increased possibility of wick dry-out and LHP deprime.


Chapter 3: LHP Design

3.1 Problem Statement

As previously discussed, one of the main objectives of this study was to design and

manufacture a functional loop heat pipe. The purpose of developing a LHP was two

fold: to demonstrate LHPs as a reliable and robust technology; and to use the results as a

stepping stone for the development of much smaller and more efficient systems.

Therefore, no rigid performance or property constraints were set; however, maximizing

the heat load limit while also minimizing the mass and dimensions of the LHP were a

priority. As will be shown, primary components were selected based on their relative

costs and availability as well as their performance in previous tests found in literature.

Below is a set of parameters that, wherever possible, were employed:

• Maximum Heat Load: 20 to 100W

• Temperature range: 10 to 80 °C

• Maximum diameter of evaporator: 25.4 mm

• Maximum length of active evaporator 150 mm

• Maximum diameter of transportation lines: 10mm

• Heat transfer distance: 200 to 500 mm

44


CHAPTER 3: LHP DESIGN 45

• Condenser length: 100 to 200 mm

• Cooling conditions: liquid cooling

3.2 Design Approach

A total of three LHPs were designed, constructed, and tested in sequence. The first LHP

was designed using a similar procedure to the one described in Section 2.1. Preliminary

working fluids, wick materials, and case materials were chosen based on flight heritage,

cost, availability, and performance characteristics. Mathematical models were then used

to estimate working fluid properties, reservoir volume, fluid inventory, overall mass,

pressure losses, and maximum heat load. Based on these results, transportation lines

were optimized to minimize pressure losses and therefore maximize the potential heat

load limit. Similarly, the wick structure was designed to maximize efficiency of fluid

flow through the wick. Material selection was then finalized based primarily on

compatibility and method of loop sealing. Each subsequent LHP was developed with

only minor modifications to address any performance issues of the previous unit tested.

The following sections summarize the design and selection process. A brief overview of

the final sizes and properties of all three LHP’s is presented in section 3.10.

3.3 Wick Material Selection

Two wick materials were considered for this study: ultra high molecular weight

(UHMW) polyethylene tubing and copper mesh. UHMW polyethylene tubing has been

used considerably and is well documented as a wicking material for capillary pumped



loops. Recently, polyethylene has also been applied and tested for use in LHPs with

success. For example, Khrustalev (2001) reported manufacturing an inexpensive

ammonia LHP with a UHMW polyethylene wick operated in vertical and horizontal

orientations transporting 200 W with a transport length of about one meter. The

operating temperature range was from 20 °C to 50 °C. Polyethylene is an excellent

wicking material due to its low thermal conductivity, machineability and compatibility to

most working fluids and all case materials. It is also widely available as a commercial

off the shelf (COTS) product in bulk quantities thus reducing procurement time and costs.

One drawback, however, is that the effective pore size and porosity is relatively large

(15-20 microns and 20-35% respectively) in comparison to established sintered metal

wicks. More desirable properties can be obtained but with significant increases to tooling

costs. Copper mesh has also been well documented in literature. However, its use has

been limited to conventional heat pipes due to very large pore sizes. The surface pore

size of mesh wicks is inversely proportional to the mesh number, which is defined as the

number of pores per inch (Chi, 1976). For example, a typical mesh number of 100

corresponds to a pore size of 0.01 inches or 254 microns. Therefore, it was not a viable

option for use as a primary wick but was considered for use as a secondary wick. It

should be noted that sintered metals and ceramics were initially considered. Sintered

nickel, titanium, and copper wicks have been used predominantly as a primary wicking

material in LHPs because of the ability to manufacture excellent properties (<1 /.un pore

size and >50% porosity). Unfortunately, costs for these materials are quite high and



suppliers are limited. Similarly, ceramics have the potential to be excellent wicking

materials with very low pore sizes, relatively high porosity, and very low thermal

conductivity. Unfortunately, there has not been a lot of research conducted in this area

and suppliers are also scarce. Therefore, of the two remaining choices, UHMW

polyethylene was selected as the primary wick material mainly due to its availability and

extensive use in CPLs. Two polyethylene tubes were obtained and rated by the

manufacturer to have an effective pore size of ~ 20 /um and porosity of approximately

35%. The first tube was 36 in (910 mm) long with a 1 in (25.4 mm) outer diameter and a

Vi in (12.7 mm) inner diameter while the second tube was 36 in (910 mm) long with a Vi

in (12.7 mm) outer diameter and a % in (6.35 mm) inner diameter. The material density

was estimated to be 965 kg/m3 while the material conductivity was approximately at 0.45

W/mK.

3.4 Working Fluid Selection

As specified in the problem statement, the operational temperature range of the LHP

should be approximately 10 to 80 °C. Potential working fluids that meet this requirement

and were considered for the study include acetone, methanol, and water. Details of each

fluid are provided in Table 3-1.



Table 3-1: Potential Working Fluids and their Properties1

SurfaceTension

(N/m)

LatentHeat

(J/kg)

LiquidDensity(kg/m3)

LiquidViscosity

(m2/s2)Acetone 0.0227757 530450 787.847 0.000322475Methanol 0.0233655 1190470 794.63 0.000591358

Water 0.0727734 2443540 995.683 0.00100148Ammonia 0.0212896 1182620 604.834 0.000146047

Traditionally, ammonia has been used for many space and electronics applications due to

its wide operating temperature range and outstanding merit number in comparison to

other working fluids. However, due to its relative high vapour pressure and toxicity at

room temperature, it was not considered for this application. Instead, it was used as a

baseline for performance. Water has a very high latent heat and surface tension

compared to the other fluids, and is non toxic. Its advantages manifest themselves at

higher temperatures and reach a maximum at 100 to 150 °C (Maidanik et al., 2005).

However, at elevated temperatures, water has a tendency to generate NCGs in the

presence of oxides with the most common case materials: aluminum and stainless steel.

The selection of UHMW polyethylene as the primary wicking material produces another

compatibility issue since it is hydrophobic. Polyethylene can be rendered hydrophilic

through a chemical process; however, such a method may produce NCGs within the loop.

Since the required heat dissipation is relatively low, other fluids such as acetone and

methanol were other interesting options. They present several advantages such as sub-

1 All working fluid properties evaluated at a temperature of 20°C



atmospheric working pressure for operation temperatures from -60 to 80 °C. They also

have reduced handling risks, are less expensive, and unlike water, reduce the probability

of freezing in space conditions. Acetone has been well documented in literature for use

in LHPs. Some studies have reported results comparing the performance of acetone and

ammonia. For low power applications (up to 70 W), Riehl et al. (2005) verified that both

working fluids present close performance characteristics, despite a higher superheat

(temperature difference between the reservoir and evaporator) obtained when acetone

was used as the working fluid. Tests have also shown stable long-term operation

between acetone and stainless steel for a wide range of applied power. For these reasons,

acetone was selected as the primary working fluid. Methanol was selected as an

alternative working fluid to compare performance characteristics against acetone for the

same LHP design.

3.5 Case Material Selection

Various factors including the operating temperature range of the proposed device,

compatibility with the working fluid, evaporator and condenser sizes, and internal

operating pressure were considered during the selection of the case material. Internal

operating pressure was not as great of a concern since potential working fluids, as

described in Section 3.3, have sufficiently low vapour pressures between the anticipated

loop operating temperatures of 10 to 80 °C. As a result, aluminum and stainless steel

(316L) were selected as potential case materials. Due to its high thermal conductivity

(237 W/mK), aluminum reduces the thermal resistance between the heat source and wick



material thus increasing the LHP efficiency. However, high thermal conductivity also

yields high parasitic heating of the reservoir which in turn increases the loop operating

temperature. Aluminum also has a much lower density (2700 kg/m3) compared to

stainless steel (7800 kg/m3) thus reducing the overall mass. In comparison, stainless steel

has been used much more frequently in literature. It is also more responsive to different

kinds of welding techniques and possesses high strength (tensile strength of

approximately 485 MPa). Stainless steel is also compatible with a wider range of

working fluids. Specifically, stainless steel has been successfully demonstrated with

working fluids such as water, acetone, ammonia, and methanol. Aluminum is not

recommended for use with methanol and is incompatible with water due to generation of

non-condensible gases (Peterson, 1994).

3.6 Loop Sealing

Loop sealing was also of primary concern during the design phase of this study since this

process can potentially determine success or failure of the LHP. A review of several

techniques traditionally used for welding highly conductive materials was undertaken. It

was found that electron beam (EB) welding and tungsten inert gas (TIG) welding are the

most commonly used processes for stainless steel and aluminum respectively. However,

by permanently sealing the loop, it would be difficult to perform experimental studies

such as analyzing the effect of fluid inventory and wick design on loop performance. By

welding the loop for sealing, construction of a new loop would be required for each fluid

charge and wick design. Welding also increases the likelihood of exceeding the



maximum allowable temperature of UHMW polyethylene thus greatly increasing the

complexity of sealing the loop. Therefore, an alternative means of sealing the loop was

selected. Swagelok tube fittings were used to provide an efficient and time savings way

of connecting each section of the loop while a vacuum-tight valve was used to completely

seal the loop after fluid charging. This approach facilitates testing of several fluid

charges while also allowing easy insertion and removal of the wick structure. However,

in order to maintain the vacuum in the loop after assembly, this approach confines

selection of the case material to stainless steel. Ideally, this approach reduces the time

required for assembly.

3.7 Component Sizing

As a starting point the evaporator, condenser, liquid line, and vapour line sizes were

selected based on values found in literature for use in LHPs with similar requirements.

Then, based on the specific requirements defined in the problem statement, sizes were

further constrained. Tolerance-matching with the wick material and compatibility with

the Swagelok fittings were also considered. Sizes for each component were finalized

after analyzing the predicted performance based on mathematical models for a given

design.

To ease machining and insertion of the wick, a sufficiently large evaporator was

required. Also, due to the low tolerance of the polyethylene tubes, the inner diameter of

the stainless steel tubing was required to be slightly smaller than the outer diameter of the

wick material. Accordingly, standard stainless steel tubing of V2 in (12.7 mm) and 1 in



(25.4 mm) outer diameters with lengths ranging from 50 to 130 mm were selected for

analysis. Initial sizing of the transportation lines was determined relative to the size of

the evaporator. Standard stainless steel tubing ranging from 150 to 300 mm in length

and%,X and X inches (3.18, 6.35, and 9.53 mm) in diameter was investigated. As

discussed in the problem statement, the overall size of the LHP had to be minimized. To

meet this requirement and minimize the mass of the tube fittings, stainless steel tubing

with similar dimensions to that of the vapour and liquid transportation lines were selected

for the condenser. The compensation chamber is essentially an extension of the

evaporator, sharing the same material properties and radial dimensions. It is therefore

thermally and mechanically connected to the evaporator. Its specific volume was

estimated concurrently with the fluid inventory and was based on the volumes of all other

LHP components.

3.8 Wick Structure Design

Loop heat pipe performance greatly depends on the wick structure and evaporator design.

Modifying various wick characteristics such as length, diameter, thickness, number of

axial grooves, and number of circumferential grooves can result in differences in heating

area (interface between wick and evaporator body), heat transfer coefficient across the

heating area, and vapour pressure drop in the vapour channel (Chuang, 2003).

Traditionally, the primary wick consists of axial grooves extruded down the length of the

body, similar to that shown in Figure 3-1, to provide vapour flow to the vapour line.



Circumferential grooves are threaded around the body to provide vapour flow to the main

axial vapour channels as shown in Figure 3-1.

Similar to component selection, initial wick designs were based on existing systems

with proven success. A total of four axial grooves and four circumferential grooves were

chosen. Specific dimensions of each as well as the thickness between the inner and outer

diameter of the wick were optimized to minimize the pressure drop of the fluid flow

(both liquid and vapour) while also maximizing the heat transfer capacity. Since it was

determined that polyethylene would be used as the primary wicking material, vapour

grooves were machined directly on the wick structure instead of the evaporator tubing. A

solid model of the wick structure is shown in Figure 3-2. The 1 in (25.4 mm) OD plastic

tubing was chosen due to the size of the evaporator case. For a simplified assembly

process and reduced manufacturing costs, it was decided that no secondary wick was

required. It is noted that such a design may degrade LHP performance in certain testing

conditions such as adverse orientations.



Circumferential Grooves

Figure 3-1: Design of Evaporator and Primary Wick Interface

Circumferential

- ^ A x i a l Grooves

Figure 3-2: Solid Model of Wick Structure with Vapour Grooves

3.9 Mathematical Modeling

Mathematical models developed by the Russian company TAIS Ltd. were initially used

to help size components, such as the transportation lines and wick structure, by



estimating the system pressure drop and maximum heat flux. The models were available

in a bundled demo software package called EASY 2000. The package consists of four

blocks:

1. Data input (creation, reading, editing, recording, printing);

2. Calculation of working fluid parameters such as latent heat, surface tension, liquid

and vapour viscosity.. .etc;

3. Calculation of LHP parameters such as fluid inventory, reservoir size, and

component masses;

4. Calculation of LHP processes such as maximum heat flux and steady state loop

operation temperature.

The EASY 2000 demo was bundled with three applications which were used in the

development of the first two LHPs. The first program, Idprep, allows the user to input all

physical properties such as length, diameter, pore size, porosity etc. Conditions such as

ambient temperature and sink temperature for all major junctions are also available for

input. Figure 3-3 illustrates a sample of the graphical user interface (GUI) with the

properties and dimensions of an iteration of the first LHP design.



Loop Dim. a n d P iu p . Input Data

Wick Evaporotor CondenserI WIckMetalDens 9 .6 5 E + 2 EvHeatAreaLength 8.1 E-2 CoWallThickness 8 .89E -4

Wick 0 D 2 .2225E -2 EvapOutNumber 1.0E-K] CondLength

Wick I.D 1 .27E -2 EvapWallThickness 1.5875E -3 CondlJ).

WickPoreRadius CirOrNumb/Length 4.0E +1 Hydr .Resist Coef 1.0E-H3

WickPerme ability 5 .87755E -1 : CircGroovDepth 1.5E-3 CondType 1 .OE+01 WickPorosity 3.0E-1 I------CircGroovWidth

xirctjroovLanavvjcJin1.778E -3 1 InsTubeExtDia

ij WtakBoret,ength 1 .00413E-1 2 .3 2 2 2 E-2 J NumbCondSect 1 .OE+0

Artck Un-BoredLength 1 .5875E -3 Vapor Line Liquid LineWickType 1 .0E-H3 3 VLWallThickness 8 .89E -4 LLWallThickness 8 .89E -4

FlatWickWidth VLLength 2.94E-1 LLLength 2.94E-1FlatWickThickne.ss VLHydrDia LLHydrDia 4 .5 7 2 E -3

NumbA/Gtooyesi

4.0E-H3 Hydr. Resist. Coet. 1.0E + 0 LLNumb 1 .OE+01 AxGrooveDepth 1 .5E -3 NumbVLs 1 .0E + 0 AddLLLength 0.0E-H3

AAGrooveWidth 6 0E -3

1 A • • jroo''eShape 1 .OE+O Conditionsfl-inks | Help 1 D one

1. W ick M etal D ensity , [k g /m A3] {rho for m ateria l, not for porous}

Figure 3-3: EASY 2000 User Interface and Input Parameters

The second application, Charge 1, calculates the mass of the working fluid to be charged,

compensation chamber volume, and mass of LHP components. The program uses the

physical properties specified in Idprep to simultaneously calculate fluid inventory and

compensation chamber volume using the technique described in Section 2.3.4. Then, by

specifying the density of the case material and wick structure, the approximate mass of

the LHP and each individual component is estimated. Figure 3-4 illustrates a sample of

the results obtained for the properties specified in Figure 3-3, using water as the working

fluid.



' ~ l C harge a n d W eight Solve

W orking Fluid is j W ate r ~

Liquid volum e c h an g e

V olum e of reservoir

Filled fluid m a s s

Void volum e coefficient

V olum e of th e LHP

V olum e of V apor Li n e s

V olum e of W ick w ithout hole

VoLof Hole (inside of W ick)

V olum e of C o n d en se r

V olum e of Ax. G rooves

V olum e of Liq. Lines

Voh of Circum. grooves

N um ber of Vap. and Liq.

0.559.4519.32

1.20

31 .89

4.83

7.1812.72

2,50

2i924.83

0.20

Loop M etal D ensity [7800.0

W ick M etal D ensity : 885.0

-

Input D ata File is: defau l t d a t

to'fn"3]

J L

R hoV ap(10)-'O1 Elr.1 iq(l 0 = 9 9 9 7 |

R hoV ap(80)=.29 RhoLiq(8Q)=971.8 |EWmA31

E vaporato rW eight of LHP U nits , | q 1:

Trniri-1 i dm o r e 0 .0

8.818BE+01

C o n d en se r 1.8083E401

V apor Line 3.4976E-r01

Liquid Line 3.4976E+01

R eservoir (like cube) | 3 .3209E+01

W ick 1.8138E+01

T o t a l 2.4689E+02

Tmax=r

K=

units v oiume, ~/a

374,0

1.2

! ->r t l i r if

r * i ca i- f t 1

lent.1 * 1

R ecalc

Print Window

Print to File

Minimize

Exit

Figure 3-4: Calculation of Fluid Inventory, Reservoir Volume, and LHP Mass

The third application, Cqmax, calculates the maximum heat flux which is determined by

the maximum capillary pressure developed by the wick and the pressure balance in the

LHP. Details of the method of calculation is provided in the EASY2000 User’s Manual

(TAIS, 2000). Figure 3-5 provides a sample calculation of the estimated pressure drop

for the properties specified in Figure 3-3, with water as the working fluid. In this

example, the maximum pressure drop occurs in the condenser line and the maximum heat

flux is estimated to be 285.9W, well above the range specified in the problem statement.

Using acetone and methanol as the working fluid will lower this value, especially at

lower temperatures. It should also be noted that the software uses many assumptions that



do not directly apply to the given design such as the use of a secondary wick and bayonet

tube.

C<il< ulfilion ol Qnifix

I s , C= 40 G user,W =

M ass Flow R a te , kg/ s =|| 1.1 S78E-04

Q M axim um , W = || 285.9Sum m . p re s s drop, P a =|| 6962.4

V apor line p re ss drop, P a =|| 1894.5

W ick pr drop, Pa=|| 146.7

W ick Axial G rooves, P a =[| 1224.5

Liquid line p ress, drop, P a =|| 2.1

C ondenser pr drop, P a =|| 3680.3

Add liquid line pr drop, P a =|| 0.0

Circum Groove p re ss dr, P a =)| 14.2

A cce le ra tio n 'p ress , dr. (rh o T fh ), P a = | 0.0

l-api lary o '? ; : .•••' im a -i P c r || 6957.7

%

Pressure Drops, %

100

806040200

Select Q or i s and Press "Enter'1 for Solve. P lease Done Print W indow

Figure 3-5: Pressure Balance and Maximum Heat Flux Estimate for Given Design

Unfortunately, the demo version of EASY 2000 was limited in many aspects;

working fluid selections were restricted to water and ammonia while documentation of

mathematical methods and examples of experimental validation were insufficient. It is

therefore noted here that while EASY 2000 was helpful in optimizing the sizes of each

LHP component, models with increased flexibility and functionality were required to

provide more accurate predictions of fluid inventory, compensation chamber sizing,

working fluid properties as well as operational characteristics such as steady-state

temperatures, system pressure drop and mass flow rate for any given design. As a result,



‘in-house’ software was developed in parallel with the final LHP design. The models

were developed in Matlab and are reported in Chapter 6.

3.10 Final LHP Design and Layout

After optimizing the sizes and properties of each component, the final design was

selected. It is again emphasized that each LHP was designed, manufactured, and tested

sequentially. That is, after successfully manufacturing and testing the first LHP, two

subsequent designs were created with small variations to increase performance or to

address any issues of the previous design. Design changes for each LHP are addressed in

Chapter 5, where test results of each unit are also presented and analyzed. A summary of

all properties and dimensions of the three designs is provided in

Table 3-2.

Table 3-2: Detailed Properties of each Test Unit

LHP 1 LHP 2 LHP 3Working Fluid Acetone Acetone, Methanol Acetone

Designed Temperature Range 5 to 75 °C 5 to 75 °C 5 to 75 °C

Evaporator Stainless Steel Stainless Steel Stainless SteelHeated Length 8.1 cm 8.1 cm 8.1 cmWall Thickness 1.59 mm 1.59 mm 2.11 mm

Primary Wick UHMW Polyethylene UHMW Polyethylene UHMW PolyethyleneDensity 965 m3 965 m3 965 m3Outer Diameter 2.2225 cm 2.2225 cm 2.12 cmInner Diameter 1.27 cm 1.27 cm 1.27 cmLength 10 cm 18 cm 9cm

Effective Pore Radius 20 |J m 20 |J m 20 |J m

Permeability 5.88E-13 m2 5.88E-13 m2 5.88E-13 m2



Porosity 0.35 0.35 0.35

Number of Axial Grooves 4 4 4

Number of Circ. Grooves 4 4 4

Compensation Chamber Stainless Steel Stainless Steel Stainless SteelInner Diameter 2.2225 cm 1.91 cm 2.2225 cm

Length 4.25 cm 6 cm 5.25 cmVapour Line Stainless Steel Stainless Steel Stainless Steel

Line Inner Diameter 4.6 mm 4.6 mm 4.6 mm

Line Length 29.4 cm 40.24 cm 26.4 cm

Wall Thickness 0.9 mm 0.9 mm 0.9 mm

Liquid Line Stainless Steel Stainless Steel Stainless SteelLine Inner Diameter 4.6 mm 4.6 mm 4.6 mmLine Length 29.4 cm 36.2 cm 26.4 cm

Wall Thickness 0.9 mm 0.9 mm 0.9 mmCondenser Stainless Steel Stainless Steel Stainless Steel

Line Inner Diameter 4.6 mm 4.6 mm 4.6 mmLine Length 15.2 cm 26.2 cm 15.2 cm

Wall Thickness 0.9 mm 0.9 mm 0.9 mm


Chapter 4: LHP Manufacturing

4.1 Fabrication

A loop heat pipe is essentially composed of four basic components; working fluid, wick

structure, case materials, and a sealing mechanism. Selection and design of each

component was discussed in detail in the previous chapter. In the following chapter, LHP

manufacturing techniques are discussed. Figure 4-1 illustrates the basic operations

involved in manufacturing. The basic elements, as can be seen from the figure, are parts

manufacturing, cleaning, assembly, evacuation and charging, sealing, and

validation/acceptance testing.

61


CHAPTER 4: LHP MANUFACTURING 62

Clean

Insert Wick in Envelop

Purify and Measure Fluid

Seal Loop

Heat Treat

Component Assembly

Cut and Machine Case Materials

Clean and Machine Wick

Evacuate and Charge

ACCEPTANCETESTING

Figure 4-1: Loop Heat Pipe Manufacturing Process Adapted from (Chi, 1976)

4.1.1 Case Materials

The first step in fabrication involved cutting both 1 in (25.4 mm) and 'A in (6.35 mm)

stainless steel tubing to the required lengths for each section of the LHP. The tubing was

first cut with a band saw, as shown in Figure 4-2, to the approximate lengths required

plus some small margin. Then, using a lathe, shown in Figure 4-3, both ends of each tube

were squared off and machined to a more accurate length. Care was taken to avoid



distorting the tube ends to ensure a vacuum-tight fit with the Swagelok unions during

assembly. The lathe was also used to debur the inner and outer edges at each end of the

tubes to safeguard against any debris that may damage LHP operation and to allow for

easy insertion of the wick.

Figure 4-2: Band Saw used to Cut Stainless Steel Tubing

Figure 4-3: Lathe used to Round-Off Edges of Tubing

It should be noted that the 1 in (25.4 mm) tubing, used for the evaporator casing, was cut

to a length 2.5 cm greater than the length of the reservoir and wick combined. This is due

to the fact that the Swagelok reducers, used to join the evaporator tubing to the vapour



and liquid lines, create a slight bend around the circumference of the tubing and may

damage the wick or reduce wick-wall contact which should be avoided. Further details

of the Swagelok unions, reducers, and valves used for assembly are presented in Section

4.3.1.

4.1.2 Wick

The next step in fabrication involved machining the wick structure. As discussed in

Section 3.3, a UHMW polyethylene wicking material with a 1 in (22.23 mm) outer

diameter and a Vi in (12.7 mm) inner diameter was obtained. Since the diameter of the

plastic tubing was larger than that of the stainless steel body, the outer diameter of the

wick was machined to reduce its dimensions and allow for insertion of the wick. It is

noted that the wick structure was machined to a diameter slightly larger than the internal

diameter of the stainless steel case to ensure constant wick contact with the evaporator

wall (throughout the entire length of the wick) after insertion. Since the resistance

between the heat pipe case material and the wicking structure can make up a significant

portion of the overall LHP thermal resistance, it is necessary to maintain good contact

between the two. A lathe machine was used to reduce the outer diameter of the wick, as

shown in Figure 4-4. The lathe was also used to machine four circumferential grooves on

the wick. A bit with the required width of 1.5 mm was placed at intervals of 2 cm along

the length of the wick and machined to a depth of 1.5 mm thus reducing the diameter at

each groove to 19.2 cm. Finally, the lathe was used to machine the wick to the required

length and round off any edges. Care was taken to minimize any damage or



contamination to the wick from metal filings that may have been left over from the

previous user of the machine.

Figure 4-4: Lathe used to Reduce Diameter and Machine Four Circumferential Grooves

Next, a milling machine was used to mill four axial grooves, shown in Figure 4-5. The

grooves were 6 mm in width and 1.5 mm in depth and were machined along the length of

the tubing. At one end, a small section 5 mm long was left untouched to stop vapour

flow to the compensation chamber. At the opposite end, as shown in Figure 4-6, a plastic

cap was installed to prevent liquid from flowing to the vapour line. The cap consisted of

polyethylene filler that was sintered to existing wick material. Care was taken to ensure

that the filler did not interfere with any of the vapour grooves. The depth of the filler was

limited to 2 mm to minimize damage to surrounding pores thus maximizing the active

length of the wick.



Figure 4-5: Mill used to Machine Four Wick Axial Grooves

M achined P o ly e th y len e T ubing

P la stic C ap

Figure 4-6: Wick Plastic Cap

4.1.3 Heat Source and Heat Sink Saddles

The evaporator saddle consisted of an aluminum base plate with approximate dimensions

of 75 mm by 50 mm by 20 mm and a stainless steel base plate with dimensions of 75 mm

by 50 mm by 3 mm. As shown in Figure 4-7, a Vi in (12.7 mm) semi-circular groove was

machined along the length of the base plate. A drill press, also shown in Figure 4-7, was

then used to machine four holes through the cover plate and halfway through the base



plate. Then, using a hole tap shown Figure 4-8, 10-32 threads were introduced in the

base plate. The final assembly, shown in Figure 4-9 required four screws torqued to 45

in-lb (5 N-m). Similar to the evaporator saddle, the first heat sink and subcooler saddle

was constructed by milling a lA in (6.35 mm) semi-circular groove across the length of a

150 mm by 50 mm by 6.35 mm copper plate. A total of 10 holes were machined through

both the base plate and cover plate. The final assembly of each saddle, as shown in

Figure 4-10, required ten screws torqued to 45 in-lb (5 N-m).

Figure 4-7: Milling Machine and Drill Press used for Manufacturing Saddles

Figure 4-8: Threads Machined using a 10-32 Hole Tap



4X Screws

Steel Cover Plate

Aluminum Base

Figure 4-9: Final Evaporator Saddle Assembly

10X Screws (condenser) Steel Cover Plate Condenser Saddle

Subcooler Saddle

Copper Base Plate

Cooling Line

Figure 4-10: Final Condenser and Subcooler Saddle Assembly

A second heat sink was designed and manufactured using acrylic tubing, as shown in

Figure 4-11. First, 1” (25.4 mm) and Vi” (12.7 mm) plastic tubing was cut using a band

saw to the appropriate lengths. Then, similar to the techniques used for machining the

stainless steel transportation lines, a lathe was used to round off the ends to help secure a

liquid-tight seal at the ends. Plastic adhesive was used to fasten the end caps and



inlet/outlet tubing to the main body. Finally, caulking was applied between the end cap

and steel tubing to avoid leaking of the cooling fluid.

Adhesive

io HCaulking

W V *ill

’ End Cap js’

1 Inlet Tubing

Outlet Tubing

Figure 4-11: Tube-in-Tube Parallel Flow Heat Exchanger

4.2 Component Cleaning

As previously discussed, the presence of contaminants within the LHP may degrade

performance over time. Therefore, thorough cleaning of all components before assembly

is required. The particular cleaning procedure employed for a given LHP is greatly

dependent upon the case material, wicking structure, and working fluid combination

being used. There are numerous references available that describe various cleaning

procedures for a wide variety of combinations ((Chi, 1976) and (Peterson, 1994)). The

basic steps involve the following: initial cleaning to remove any debris, metal

fillings...etc; chemical cleaning process to remove any water, oils, or films on the case or

wick materials; a series a final rinse processes to remove any remaining solvent; and in

some cases a vacuum bakeout. The best methods for cleaning stainless steels use

passivation techniques to remove or eliminate any steel particles embedded in or near the



surface from the machining process (Peterson, 1994). In addition, common passivation

treatments including nitric acid (HNO3) solutions or pastes enhance the formation of the

protective film thus increasing resistance to corrosion.

The following procedure was used to clean the stainless steel tubes (adapted from

(Peterson, 1994)):

1. Soak and brush the surfaces with a wire brush and acetone to remove oil and dirt.

2. Flush parts with acetone.

3. Soak in a solution of 15% nitric acid, 5% hydrochloric acid, and 80% water for 15

minutes.

4. Soak in a 15% nitric acid, 85% water solution at 65 °C for 15 minutes to dissolve

any iron or steel particles.

5. Flush parts with acetone or isopropyl alcohol.

6. Dry thoroughly with clean, dry air.

This procedure is compatible with both acetone and methanol working fluids. Acetone

was used to clean the wick structure of any remaining contaminants left over from

machining. A bakeout was also performed on the LHP, and is discussed in the

evacuation and charging procedures.

4.3 Assembly Procedure

Assembly of each LHP component should occur immediately after cleaning, where

applicable. Typically, assembly of loop heat pipe parts includes welding of the end cap,

welding of the case sections and wick insertion. For this particular study, the wick



structure was first inserted, as shown in Figure 4-12, into the evaporator section

immediately after uniformly heating the evaporator to a temperature of 80 °C. After

insertion, the evaporator-wick assembly was visually inspected to confirm that the wick

was in direct contact with the evaporator wall, as shown in Figure 4-13. The evaporator

and transportation lines were then connected via Swagelok unions and the loop was

sealed with a Swagelok valve. After assembly, the apparatus was weighed for

comparison to the mass after fluid charging.

Figure 4-12: Wick-Evaporator Insertion

Wick Structure

Evaporator Wall Contact

Figure 4-13: Evaporator Wall-Wick Contact after Insertion



4.3.1 Swagelok Fittings

Swagelok fittings were chosen as an easy and efficient way of assembling and sealing

LHP components. The fittings use a two ferrule design that provides leak-tight seals up

to a pressure of 10'7 Torr (1.33x10 5Pa). The two ferrule design separates sealing and

tube retaining functions. As shown in Figure 4-14, the front ferrule creates a seal against

the fitting body and tube OD, while the back ferrule provides a radial tube grip that can

withstand vibrations and temperature variations.

Front Ferrule Seal

Back Ferrule Grip

Figure 4-14: Swagelok Two-Ferrule Design (Swagelok Tube Fittings, 2004)

Nine fittings were used for assembly of the first and third LHP and are illustrated in

Figure 4-15. Two reducing unions (1 in (25.4 mm) OD to lA in (6.35 mm) OD) were

used to connect the evaporator section to both the vapour and liquid lines. A union tee

(% in (6.35 mm) OD) was used to connect the liquid line to the fill-tube and sealing

valve. A Bellows valve was used for fluid charging and sealing the system. Three union

elbows ( ]/4 in (6.35 mm) OD) were installed at three of the four right angles of the loop.

The second LHP was assembled in the same manner as LHP 1 and 3 with the exception



of the top two union elbows. These two fittings were replaced with 90° degree bends in

the tubing. The purpose of replacing the fittings was to reduce pressure losses introduced

by the fittings.

Elbow Fittings

Bellows Valve

Port Connectors

Union Tee 1" to y* ' Reducers

Figure 4-15: Swagelok Fittings

The following are installation procedures of all Swagelok fittings under 1 in OD:

1. Insert tubing into the Swagelok tube fitting.

2. Make sure that the stainless steel tubing rests firmly on the shoulder of the tube

fitting body and that the nut is finger-tight.

3. Scribe the nut at the 6 o’clock position.

4. While holding the fitting body steady, tighten the nut 1 and lA turns to the 9

o’clock position.



One of the main advantages of using Swagelok fittings is that each fitting may be

disassembled and reassembled several times. This function permits easy access to the

wick structure for inspection. It also enables testing of varying fluid inventories in the

same loop without the requirement of manufacturing a new prototype for every charge.

Reassembly instructions are as follows: first, insert the tubing with preswaged ferrules

into the fitting body until the front ferrule sits flush with the fitting, then rotate the nut

with a wrench to the previously pulled-up position, at which point a significant increase

in resistance will be encountered; tighten slightly with a wrench.

It is noted that in some instances, port connectors were used as a substitute to regular

stainless steel tubing in areas where only a small length of tubing was required. The

location of each port connector used in LHP 1 and 3 is shown in Figure 4-15, while

Figure 4-16 illustrates the component before assembly. The tube end of the port

connectors are installed the same way as regular stainless steel tubing. On the machined

ferrule end, the procedures are as follows: remove the nut and ferrules from one the tube

fittings; then, place the nut of the machined ferrule end of the port connector; turn the nut

onto the fitting so that it is finger tight; scribe the nut at the 6 o’clock position; while

holding fitting body steady, tighten the nut lA of a turn to the 9 o’clock position.



Tube End

Machine Ferrule

Figure 4-16: Port Connector

4.4 Evacuation and Charging Procedures

Particular attention to loop evacuation and fluid charging is necessary as it is a critical

stage in LHP manufacturing. The requirement of evacuating the LHP relates to the

evaporation process of the working fluid. Based on the ideal gas law expressed in Eq.

(4.1), temperature of a gas is directly proportional to its pressure.

p V = nRT (4.1)

where P is the absolute pressure; V is the volume; n is the number of moles; R is the

universal gas constant = 8.3145 J/mol K; T is the absolute temperature. Therefore, if the

pressure in a constant volume is decreased then the temperature at which evaporation

occurs in a liquid is also decreased. For example, at atmospheric pressure (101.33 kPa),

water boils at approximately 100 °C. If the pressure is reduced to 3.17 kPa, water will

instead boil at approximately 25 °C. Similarly, acetone and methanol both boil at 25 °C

for pressures of 26.66 kPa and 16.85 kPa respectively. Thus, start-up at low temperatures

is partly established by reducing the pressure in the LHP.



Another important factor for evacuation is the need to remove any remaining

particulates that may damage LHP operation. More specifically, a LHP should be

evacuated and heated prior to charging to remove any foreign materials that may

subsequently appear as unwanted non-condensibles, or that chemically react with the

working fluid forming undesirable corrosion products. A general rule is to evacuate the

pipe at a temperature greater than the heat pipe operating temperature (Chi, 1976). This

process is sometimes referred to as bakeout.

Figure 4-17 illustrates the apparatus used for evacuation and charging. The system is

composed of a six primary components: vacuum pump, trap, flow lines, pressure gauge,

valves, and fluid container.

Fluid C o n t a i n e rP r e s s u r e G a u g e

L H P A s s e m b l y

Figure 4-17: Evacuation and Charging System Assembly



The vaccum pump is an Alcatel SD series rotary vane pump. It has a nominal flow rate

of 6.5 m3/h or 1.8 L/s and produces a minimum pressure of 1.5xlO“3Torr (0.2 Pa),

suitable for many LHP applications. The flow line system consisted of lA in (6.35 mm)

stainless steel tubing connected with Swagelok fittings similar to those used in assembly

of each LHP. A trap was installed at the pump inlet to help prevent entry of pump oils

and condensable gases. A Kurt J. Lesker pressure gauge was installed directly after the

trap, between the vacuum pump and the LHP, and has a minimum pressure reading of 10'

3 Torr (0.133 Pa). Three Swagelok Bellows valves were installed in the indicated

location, shown in Figure 4-17. Valve 1 controls the main line to the vacuum pump. It is

opened to connect the vacuum pump to the LHP for evacuation. During fluid charging,

valve 1 is closed to halt any fluid from entering and damaging the vacuum pump, trap, or

pressure gauge. Valve 2 controls the flow of the working fluid and is therefore closed

during the entire evacuation process. It is only slightly opened while charging the

necessary amount of fluid to the LHP. Valve 3 is first installed on the LHP and then

placed directly below the fluid flow line for access to both the vacuum and fluid charge

lines. It remains open during evacuation and fluid charging procedures and is closed for

final sealing of the LHP. The fluid container provides the required amount of working

fluid. It is essentially a graduated cylinder permanently installed on the end of valve 2.

A summary of the standard evacuation and charging procedures are provided below:

1. Assemble the LHP and attach to valve 3.

2. Seal valve 3 and attach to the vacuum assembly.



3. Fill the fluid container with the required volume of working fluid (at ambient

temperature); ensure no air bubbles are trapped in the valve.

4. Seal valve 2 and attach to the vacuum assembly.

5. Confirm valves 1, 2, and 3 are closed and then turn on the vacuum pump.

6. Once the pressure gauge has stabilized, open valve 1.

7. Let the system stabilize for 20 minutes then check the pressure reading; the

pressure gauge should read a maximum of 10 mTorr. If this is not the case, then

the system may contain some residual liquid in the flow lines. The entire

assembly must then be heated to an elevated temperature to boil off the remaining

liquid. As the lines are heated there should be a sudden jump in the pressure

reading, indicating vaporization of the residual liquid.

8. After heating, let the system cool to ambient temperature. It may be necessary to

repeat this process several times over the course of several hours until the

appropriate pressure reading is obtained.

9. Open valve 3 to evacuate the LHP. The loop assembly may also contain residual

liquid left over from the cleaning process. In such a case, repeat the same process

described in Steps 7 and 8 until the required pressure is obtained. Care should be

taken to avoid heating the evaporator to significant levels since the wick structure

may be damaged at temperatures above 80 °C.

10. Once the pressure has stabilized at approximately 3 to 5 mTorr, close valve 3 and

turn off the pump. Record the pressure reading and leave system as is over night.



11. Repeat steps 5 through 8. Open valve 3 while watching the pressure gauge. The

pressure should be maintained at the same pressure recorded the day before. If

this is the case, then proceed to the next step. If this is not the case then there may

be a leak in the loop assembly. Remove the loop assembly from the system and

check all Swagelok connections. Repeat steps 5 through 8 again.

12. Close valve 1 and make sure valve 3 is open. Slowly open valve 2 while

inspecting the fluid container. Close valve 2 when the required volume is

charged.

13. Apply a cold source, such as an ice pack, to the loop assembly to ensure that all of

the fluid is charged to the loop.

14. Close valve 3 and detach from the vacuum assembly.

15. Turn off the vacuum pump and seal all outlets.

4.5 LHP Sealing

After evacuation and fluid charging, the LHP was sealed by closing the Swagelok

Bellows valve. The apparatus was then removed from the vacuum system and a cap was

placed on the end of the valve. The LHP was then weighed to verify that the correct

amount of fluid was charged. Performance verification tests were performed and are

described in detail in the following chapter. The final LHP assembly is shown below in

Figure 4-18.



Condenser j

VapourLine

ompensation f ' , Chamber . -i.'S0

Evaporator

Figure 4-18: Final LHP Assembly


Chapter 5: LHP Testing

5.1 Experimental Approach

Preliminary LHP testing of each apparatus was required as part of the evaluation

procedure necessary for selecting an optimum design that meets the problem

specifications. All three LHPs were tested in succession to verify performance and stable

operation. The first phase of testing focused on loop start-up and steady-state operating

temperatures in comparison to ‘baseline’ results obtained from each LHP without

working fluid. After testing each unit, design changes were implemented in each

subsequent LHP to increase performance, with the final LHP producing the best results.

As a result, the final LHP was selected for further analysis. The second phase of testing

studied the effect of fluid inventory on LHP start-up, steady-state operating temperature,

effective thermal resistance, and heat transfer coefficient. The final phase of testing was

used to measure the effect of orientation and periodic heating on thermal performance.

Temperature hysteresis was also investigated during the final phase of testing.

5.2 Test Specimen and Experimental Setup

All three test articles in this study were constructed primarily of stainless steel, with

UHMW polyethylene as the wick material. A summary of the specifications of each test

81


CHAPTER 5: LHP TESTING 82

unit is provided in Table 3-2. All tests were performed at the Carleton University Heat

Pipe Laboratory in Ottawa, Ontario. A test platform was constructed for the purpose of

LHP performance verification and characterization of properties. The overall test

platform is illustrated in Figure 5-1. A reference coordinate system fixed to the test

board, also shown in Figure 5-1, was used to describe different orientations of the LHP.

The evaporator and condenser sections were aligned with the X axis while the vapour and

liquid lines were aligned with the Y axis. The Z axis was parallel but opposite in

direction to the gravity vector (coming out of the page in Figure 5-1).

Chiller 5 - 1 0 °C

Keithley Software

X

Test Board Data Acquisition Unit

2X MultimetersO 0 Vo|fage Regulator (Variac)

Figure 5-1: Schematic of Test Platform in Heat Pipe Laboratory



5.2.1 Orientation of Test Unit

Nominal testing of each unit proceeded with a fixed elevation. That is, the liquid and

vapour lines were fixed horizontally in the y-axis. The elevation refers to the vertical

distance between the centre of the evaporator and the centre of the condenser line. The

evaporator and reservoir were also fixed in the horizontal position (in the x-axis),

producing no adverse tilt. The tilt refers to the vertical distance between the centre of the

evaporator and the centre of the reservoir.

5.2.2 Instrumentation

Temperature Sensors

A minimum of 11 Omega T-type copper/constantan thermocouples were used to monitor

LHP temperatures. Thermocouples were placed at critical points along the LHP,

including the evaporator, vapour line inlet, centre of vapour line, condenser, liquid line

inlet, centre of liquid line, and compensation chamber. Thermocouples were also used to

monitor ambient temperature, as well as the temperature of the cooling liquid entering

and exiting the condenser heat exchanger. The uncertainty of the thermocouple readings

were estimated to be approximately +/- 0.5 °C. Aluminum adhesive tape was used to fix

the thermocouples to all testing surfaces and had a maximum operating temperature of

120 °C. Figure 5-2 illustrates the location of thermocouples for two condenser

configurations used during testing.



TCC5TC11 TC10

TC4TC5

TC11

TC9 •Ambient

TC3 TC3TC6 TC6 TC9 *Ambient

TC8-TC1 C8

TC2TC7 TC1TC7

Figure 5-2: Thermocouple Placement for both Configurations

Data Acquisition System

A Keithley Model 2700 data acquisition system was employed during LHP testing and

used a built-in digital input board to read the potential from the thermocouples. The

temperatures were recorded and displayed in real time through a direct interface card to a

personal computer using Keithley Integra Series Software. All data was automatically

output as .dat files at the end of each test procedure. Matlab scripts were used for post

processing of data.

Heat Source & Power Control System

An Omega kapton flexible heater provided up to 60 W of power to the LHP evaporator.

The voltage supplied to the heater was controlled by a 110 V variac. One multimeter

monitored the voltage input to the heater while another multimeter monitored the

amperage. The resolution of the voltage reading was 0.1 V while the resolution of the



amperage reading was 0.01 A. The uncertainty for the power input was estimated to be

approximately +/- 2% of the reading.

Cooling System

A Lytron recirculation chiller was used to provide a continuous supply of cooling liquid

at a constant temperature and volumetric flow rate. During operation, a minimum sink

plate temperature of 5 °C was achieved and maintained within +/- 0.5 °C using an

internal thermostat. The cooling fluid used during operation was a mixture of 70% water

and 30% ethylene glycol. The maximum operating volume flow rate is 6.8 L/min. The

nominal volume flow rate during testing was 9 L/min.

5.2.3 Insulation and Test Frame

Nomex piping insulation (3/8” (9.5 mm) thick) was wrapped around the evaporator,

condenser, compensation chamber, and liquid and vapour lines to minimize the influence

of the environment (ie. to control parasitic heating and cooling during testing). All LHP

components were insulated separately to reduce any external heat paths between them.

This was vitally important between the evaporator and CC since heat leak to the CC must

be minimized in each design.

Two test stands were built to support the LHP for levelled testing and to minimize the

effects of small vibrations. The first test stand consisted of a large table with wooden

mounting blocks. The mounting block heights were adjusted for tilt and elevation tests.

The second test stand was constructed of steel frames and supported a conventional heat



exchanger type condenser. However, this design did not support tilt or elevation tests.

During levelled testing, both test stands ensured that the evaporator was at the same level

as the condenser within +/- 1 mm. The evaporator was also maintained at the same

elevation as the CC within +1-2 mm to minimize any gravitational effect on pressure

gradient across the wick.

5.2.4 Evaporator-Heat Source Assembly

An aluminum saddle was constructed to house to the evaporator and effectively provide

the required heat input. Figure 5-3 provides a schematic of the evaporator-heat source

assembly.

1" (25.4 mm) OD Steel Cover Plate

AluminumBase

ThermalCompoundHeating

Element

Figure 5-3: Evaporator-Heat Source Assembly

As shown in Figure 5-3, the saddle consists of a base and a cover plate. A flexible

heating element was permanently attached to the flat surface of the saddle base using a

thin and consistent layer of thermal epoxy. The selected thermal epoxy had a relatively

high thermal conductivity (1.05 W j m K ) to minimize the thermal resistance between the



heating element and saddle. The aluminum base simulated a typical thermal mass that

may be encountered during normal operation of the LHP. The relatively high thermal

conductivity of aluminum also ensured a uniform heat flux was applied to the evaporator.

The stainless steel saddle cover was also used to ensure a uniform heat flux was applied

to the evaporator. By torquing the four screws to 45 in-lb (5 N-m), the cover provided

good contact between the evaporator outer wall and the saddle base, thus increasing the

thermal conductance between the two. To maximize the thermal conductance between

the evaporator wall and saddle base and ensure a uniform temperature distribution, a

thermal joint compound (Wakefield Engineering 120 Series: 0.735 W/mK) was applied

between contact surfaces.

5.2.5 Sink Assembly

Active cooling of the condenser region during testing was necessary due to a relatively

small heat transfer area in the condenser and low operating temperature of the wick

material. As a result, two heat exchangers were constructed for the condenser. Figures

Figure 5-4 and Figure 5-5 illustrate both designs.



Steel Cover Plate

Thermal Compound

5X ScrewsCondenser W (6.35 mm) OD

i I / i '

CopperBase

Cooling Line 5/6” (21 mm) OD

Figure 5-4: Condenser-Heat Exchanger Assembly 1

Cooling Fluid In letAcrylic Tubing

Condenser % " (6.35 mm) OD

Cooling Fluid Outlet

Cooling Line 5/6" (21 mm) OD

Figure 5-5: Condenser-Heat Exchanger Assembly 2

The first sink design is similar to that of the heater assembly. The condenser tubing is

sandwiched between a copper base plate with a %” (6.35 mm) wide groove and a

stainless steel cover. To maintain good contact between the condenser and the copper

base plate, 10 screws were torqued to 45 in-lb (5 N-m). A thermal joint compound



(Wakefield Engineering 120 Series) was applied between contact surfaces to maximize

thermal conductance. To regulate the sink temperature, the Lytron chiller was used to

circulate the cooling fluid in copper lines below the base plate. These copper lines were

permanently soldered to the base plate to reduce thermal resistance.

The second sink design is a conventional tube-in-tube parallel-flow heat exchanger.

This design is compact and has a comparatively simple design. In this case, the cooling

fluid is in direct contact with the condenser tubing thus greatly increasing the heat

transfer from the condenser to the cooling fluid.

The chiller temperature was maintained at its minimum temperature of 5 °C for the

duration of each test procedure. In some cases, further subcooling of the liquid line was

required in order to perform tests over a wider range of applied powers and operating

temperatures. Therefore, a subcooler with an equivalent design to that of the first heat

exchanger was attached to the liquid line of the LHP. In this case, the cooling liquid

exiting the condenser heat exchanger was diverted to the subcooler before returning to

the chiller as illustrated in Figure 5-6.



Cooling Fluid InCooling Fluid

Diverted to Subcooler

Cooling Fluid Out

Figure 5-6: Addition of Subcooler to Liquid Line

5.3 Operating Procedure

To obtain consistent and reliable experimental results, a standard procedure was

implemented, as listed below:

1. Leave the test apparatus sit idle overnight for at least 12 hours2.

2. Adjust tilt and elevation of the apparatus to obtain the required orientation.

3. Ensure all thermocouples are properly connected

4. Adjust insulation of all critical areas of the apparatus.

5. Connect heating element power cord.



6. Turn on the computer and run Keithley 2700 Software; start collecting data.

7. Turn on the chiller and set chiller temperature to required sink temperature, 5-10

°C.

8. Adjust the cooling water volume flow rate to 9 L/min.

9. Wait until all temperatures including cooling fluid temperature, loop temperature,

and ambient temperature reach steady state. This takes approximately 20-30

minutes.

10. Adjust the variac to the desired heat load. Note: variac controls the heater

voltage. The desired heat load is obtained by adjusting the voltage based on the

readings of voltage and amperage from both multimeters; where power applied in

Watts is, P = V- A.

11. Wait until all loop temperatures reach steady-state values. That is, no more than a

0.5 °C variation over 20 minutes.

12. Adjust heat load to a desired intensity and repeat steps 11 and 12 for all required

heat loads for the current test. Record heat load and scan number at time of

adjustment.

13. Turn off heater and chiller, disconnect all power cords.

14. Terminate data acquisition software and export .dat file for post-processing.

2 To obtain consistent results, it is required that each test be started with similar initial conditions in the

evaporator. In particular, the initial state of the working fluid in the wick structure should be consistent as

explained in Section 2.2.2.



It is noted that for safety reasons, testing was terminated and all power disconnected if

the evaporator temperature exceeded 75 °C, chiller volumetric flow rate dropped below 2

L/min, or sink temperature dropped below 5 °C. These were constraints specified by the

wick and chiller manufacturers respectively.

5.4 Test Phase 1: Preliminary Testing

As discussed in Section 5.1, preliminary tests were carried out on each test unit to

investigate functionality. In particular, LHP start-up and stable operation were verified

under steady-state conditions. The following sections discuss the test results for each unit

and summarize the subsequent design changes for each.

5.4.1 Baseline Tests

Each unit was tested with no working fluid charged in the loop. These tests were

considered ‘baseline’ tests, and used to compare against results obtained from each

apparatus after fluid charging. In particular, the results were used to ascertain whether or

not loop start-up occurred, as well as to quickly establish how much energy was

transferred through LHP operation in comparison to pure conduction. Figure 5-7

presents the results obtained from testing the third LHP with no working fluid. It

illustrates the average temperature of the evaporator, vapour line, liquid line, and

compensation chamber for a power input of 5 and 15W and a sink temperature of 5 °C.



80

E va pora to r V a p o u r Line

Liquid LineC o m p e n s a t io n C h a m b e r

70

60

50

40

30

20

10

P o w e r0

20 40 60 80 100 1200 140 160Time (min)

Figure 5-7: Baseline Results for 5 and 10W Power Inputs

Since heat transfer is achieved primarily through conduction, the temperature of each

section closest to the heat source (evaporator and compensation chamber) rose steadily

without any abrupt fluctuations. Components located further away from the heat source

(vapour line and liquid line) also rose in temperature but at a much slower rate due to the

relatively low thermal conductivity of stainless steel. It should be noted that since the

steady-state temperature is a function of thermal conductivity and ambient conditions, the

ambient temperature and insulation were kept constant throughout all baseline tests to

ensure consistent results between tests. The steady-state temperature of the evaporator

for a range of power loads for each unit was compiled and is shown in Figure 5-8. The



trendline demonstrates the linear relationship between the applied power and resulting

steady-state temperature of the evaporator.

120

— 100 - - -

a.

Applied Power (W)

Figure 5-8: Steady-State Evaporator Temperatures for a Given Power Input (Baseline)

5.4.2 Test Apparatus #1

The dimensions and properties of the first LHP are provided in Table 3-2. The first set of

tests was performed to demonstrate the start-up capabilities of the LHP at low power

levels. Start-up is identified by a sudden rise in vapour line temperature as superheated

vapour is pushed from the evaporator to the vapour line, followed by a sudden drop in the

liquid line temperature as cold liquid is pushed from the condenser to the liquid line.

These features are clearly illustrated in Figure 5-9 for a 5 W start-up with a sink

temperature of 5 °C, thus demonstrating successful LHP operation of the test unit. A



slight temperature overshoot of 1.2 °C was also observed. Several start-ups at various

applied heat loads were attempted while maintaining the sink temperature at 5 °C. The

LHP demonstrated successful start-up for power inputs ranging from 2.5 W to 20 W.

The minimum power threshold for start-up was found to be approximately 2.5 W for this

particular design.

40

Evapora to r

35

30

25

20

Liquid Line

150 20 40 60 80 100 120T im e (min)

Figure 5-9: Temperature Profile for a 5 W Start-Up

In the second set of tests, power cycling was performed to analyze the ability of the

LHP to handle sudden changes in applied heat load. Tests involved successive

application of various heat loads with a sink temperature of 5 °C. As discussed in the

operating procedures, temperatures were allowed to reach steady state before applying a

new power input to the evaporator. Figure 5-10 and Figure 5-11 show the temperature



profiles of the test unit with respect to time during power cycling tests. These tests

revealed that the LHP was able to sufficiently handle abrupt changes in the applied power

without de-priming. Furthermore, no significant temperature overshoot was observed

while cycling the power.

Additional testing was performed to investigate consistent LHP operation over time.

These tests revealed significant degradation of loop performance, and after two weeks of

thorough testing, the unit completely failed. As shown in Figure 5-12, start-up did not

occur for a power input of 20 W. The temperature of both the evaporator and

compensation chamber increased rapidly and did not reach steady-state before the critical

cut-off temperature of 75 °C. Further testing of the LHP was halted and the fluid

inventory was discharged. Inspection of the wick revealed some minor damage to the

outer layer of the material. As a result, a second LHP was manufactured and assembled

with some minor modifications to the first design.



6D

Evapora to r V a p o u r Line


50

40

30

20

10

P o w e r

00 50 100 150 200 250Tim e (min)

Figure 5-10: Power Cycling (5-10-15-10 W)

E va p o ra to r V a p o u r L ine

L iq u id L ine — C o m p e n s a tio n C h a m b e r

CL

20

P o w e r

100 200 300 T im e (m in )

400 500 600

Figure 5-11: Power Cycling (2-20-2-20 W)



70

60; E va pora to r

50

40

30

Liquid Line

20

10

00 5 10 15 20 25 30 35 40 45

Tim e (min)

Figure 5-12: I HP Failure, 20 W Initial Input Power


After reviewing the results from the first LHP, it was found that the length of the

condenser limited the heat transfer capacity of the apparatus, thus also limiting the

available power range for testing. Therefore, a more efficient condenser-heat exchanger

interface was designed for the second LHP and is shown in Figure 4-11. This design

greatly increases the heat transfer coefficient between the cooling fluid and the working

fluid since the cooling fluid is in direct contact with the condenser thus minimizing the

thermal resistance between the two fluids. To accommodate the new interface, the

condenser was lengthened. Consequently, the length of the vapour line, liquid line, and

evaporator were also slightly increased, as summarized in Table 3-2. In addition to the



condenser-sink interface, the wick structure was also redesigned. The new wick extended

8 cm past the vapour grooves, spanning the length of the compensation chamber. The

purpose of the new wick design was twofold: to maintain a consistent supply of working

fluid to the primary wick even at high heat loads or adverse orientations thus maximizing

LHP efficiency; to help reduce heat leak from the evaporator to the compensation

chamber thus reducing the LHP operating temperature.

Similar to the previous device, the second LHP was initially tested to verify

functionality. The primary goal was to clearly demonstrate loop start-up.

Correspondingly, the unit was tested at power loads of 5, 10, 20, 30 and 40W.

Unfortunately, as illustrated in figure 5-13, no significant cooling in the liquid line was

visible. In most cases, the temperature of the evaporator and compensation chamber did

not reach steady-state before the maximum operating temperature of 75 °C. Steady-state

temperatures were only achieved with power loads of less than 5W. However, these

results were comparable to those derived from the baseline tests, suggesting that the

dominant mode of heat transfer was conduction. The primary mechanism for operational

failure was not immediately evident; it was therefore decided to revert to a design similar

to the first LHP. The wick structure was replaced with a wick comprised of the same

dimensions as the one used in the first test unit. Furthermore, the new condenser-sink

interface was replaced with the saddle used on the first test unit. The overall dimensions

of the LHP remained the same and the fluid inventory was adjusted accordingly.

Subsequent tests to verify performance produced the same results, suggesting that



operational failure was caused by a small leak in one of the lines at a Swagelok

connection. A leak was confirmed through vacuum testing of the apparatus.

70

60

E v a p o ra to r

50C o m p e n s a t i o nC h a m b e r

40

V a p o u r Lihe

30

20

10Liquid Line

00 10 20 30 40 50 60T im e (min)

Figure 5-13: LHP Failure, 10 W Initial Input Power


The third LHP was assembled using seamless stainless steel with higher tolerances to

help protect against small leaks at the Swagelok connections. It was fabricated with

similar physical properties to that of the first LHP, including the wick structure.

Furthermore, the condenser-heat exchanger interface used in the first LHP was again

used. A sub-cooler was added to the liquid line for additional heat transfer thus enabling

testing throughout a wider range of applied powers. Figure 4-18 demonstrates the final

design; component volumes are approximately the same as the first LHP. A summary of



the dimensions and properties is provided in Table 3-2. Preliminary test results showed

clear start-up at powers of 5, 10, 20, 30 and 40 W. Continuous and stable operation was

also observed over the same power range. The final LHP was therefore selected for the

second phase of testing.

5.5 Test Phase 2: Experimental Study of the Effect of Fluid Inventory

on LHP Performance

The primary objective of the second phase of testing was to determine the LHPs

sensitivity to fluid charge. In particular, sensitivity studies were conducted to evaluate

the effect of fluid inventory on LHP efficiency measured by start-up, steady-state

operating temperature, maximum transport capacity, effective thermal resistance, and

heat transfer coefficient. Based on the results obtained from the sensitivity studies, an

optimal fluid charge was selected and used for the final phase of testing and analysis.

5.5.1 Fluid Inventory

In order to effectively study the impact of fluid inventory on loop performance, four fluid

charges over a sufficient range were selected for analysis. Furthermore, to ensure test

results were comparable, the overall LHP design and in particular the volume of the

compensation chamber remained constant for all fluid charges. The compensation

chamber volume and fluid charges were estimated using the alpha/beta approach

described in Section 2.3.4. Each fluid charge corresponds to a specific void fraction in

the compensation chamber, a , for worst case hot conditions and a specific liquid-filled



fraction in the compensation chamber, /?, for worst case cold conditions. A summary of

all four fluid inventories is provided in Table 5-1.

Table 5-1: Summary of Fluid Charges

Fluid Inventory (g)Hot Case Void

Fraction, a

Cold Case Liquid-Filled

Fraction, f i

28 0 0.7

23 0.3 0.4

20 0.5 0.2

15 0.9 0

5.5.2 Start-Up

The first set of tests was performed to study the effect of fluid inventory on LHP start-up.

As previously discussed, start-up is characterized by the initialization of fluid circulation

in the loop and is typically identified by a sudden rise in the vapour line temperature

followed by a sudden drop in the liquid line temperature. Additionally, start-up can

sometimes be identified by a sudden drop in temperature difference between the

compensation chamber and evaporator. Start-up was attempted primarily for low power

loads in the range of 2 to 20 W for all four fluid charges. An overview of the start-up

results obtained for each fluid charge is presented below, followed by a comparative

analysis of the experimental data.

Fluid Inventory 1: 23 g

In each test, the LHP initially demonstrated characteristics of a typical start-up. Figure

5-1 illustrates a 20 W start-up. After 28 minutes of testing the vapour line experienced a



sudden rise in temperature and continued to rise until its temperature was approximately

equal to the temperature of the evaporator. Shortly thereafter, the liquid line experienced

a slight drop in temperature, indicating circulation of the working fluid within the loop.

However, its temperature was not sustained and began to rise at the same rate of the

evaporator and compensation chamber. Power to the evaporator heater was subsequently

turned off before the temperature of the evaporator rose above the maximum allowable

temperature of the wick. Power to the heater was not returned until the temperature of

the evaporator and compensation chamber dropped below 40 °C. Interestingly, only after

re-applying the power did the LHP temperatures reach steady-state. As shown in Figure

5-15, the temperature of the evaporator quickly stabilized around 50 °C after re-applying

20 W.

80C o m p e n s a t io n '^ffa'rrib'ei’ " T70

80

o 503

40

£ 30

- - .L iq u (d/ L i n e !

20

10

010 15 20 25 3 0 35 40 450 5

Time (min)

Figure 5-14: 20 W Start-Up, 24 g Fluid Inventory



BO

70Evapora to r

C o m p e n s a t io n60

C h a m b e r

50

40

$ u d d e n I n q e a s k ’ " ~~s Tti Vapbu"r"L[ne"? l l e m p e ra tu re

30

20Liquid Line

10

00 20 40 60 60 100 120 140Tim e (min)

Figure 5-15: Temperature Profile for 20 W Start-Up, 24 g Fluid Inventory

Fluid Inventory 2:15 g

Similar results were obtained for a fluid inventory of 15 g for power loads of 10 W and

greater. An example of a 20 W start-up is shown in Figure 5-16. For power loads of 5 W

or less, the evaporator and compensation chamber temperature reached steady-state

immediately following start-up, as shown in Figure 5-17. It is pointed out that in this

example there was no visible subcooling in the liquid line. The temperature in the liquid

line actually increases slightly and is slow to reach steady-state. At low powers, the mass

flow rate is slow and the liquid has a longer residence time in the liquid line, resulting in

parasitic heating from ambient and low subcooling. In comparison to the baseline results,



the evaporator temperature is visibly cooler (approximately 5 °C) for the same power

load. This suggests that there was some heat transfer due to fluid flow in the loop.

Evapora to r

V a p o u r '□ T ie " S.

LiquJ Line

BO 100 120 140 160Tim e (min)

180

Figure 5-16: Temperature Profile for 20 W Start-Up, 15 g Fluid Inventory



BO

70

60

50E vapora to rC o m p e n s a t io n

C h a m b e r 1 -----40

30 V a p o u r T i j ie

LineLiquid20

10

0100 120 14040 60 B00 20

Tim e (min)


Fluid Inventory 3: 28 g

The third fluid charge tested produced more reliable and consistent results compared to

the first two fluid charges. As shown in Figure 5-18 for a power input of 15 W, steady-

state temperatures were obtained immediately following start-up. An interesting result is

that the temperature of the evaporator was consistently 5 to 10 °C higher than the vapour

line and approximately 10 to 15 °C higher than the compensation chamber. These results

indicate a higher thermal resistance in the evaporator compared to previous tests. A

higher thermal resistance in the evaporator is usually attributed to vapour bubbles in the

wick structure. Trapped vapour bubbles in the wick can block some of the liquid return

paths resulting in less cooling at the inner surface of the evaporator section.



80

70

Evapora to r60

50

40V a p o u r Lin?

30

20Liquid Line

10

020 30 40 5 0 60 70 80 900 10

Time (min)


Fluid Inventory 4: 20g

The final fluid charge tested produced the most reliable and consistent results for start-up.

Fluid flow in both the vapour and liquid lines were clearly visible for a range of power

loads from 5 W to 45 W. Table 5-1 provides a summary of the start-up tests conducted

for a power range of 5 to 20 W. It also includes the steady-state temperature of the

evaporator after start-up. Figure 5-19 illustrates a typical 15 W start-up for a fluid

inventory of 20 g.



Table 5-2: Summary of Start-up Results for a Fluid Charge of 20 g

Applied Power (W)Average Sink Temperature

(°C)

AverageAmbient

Temperature(°C)

Time to Start-up (hr:min:s)

T max Of Evaporator Block (°C)

5 5.2 18.3 0:24:36 44.810 5.5 19.0 0:19:11 45.3

15 5.3 18.1 0:10:06 48.520 5.0 18.5 0:04:54 50.1

BO

70

60

Evapora tor50

C o m p en sa t io n C h a m b er i 1----------r - ---------40

V a p o u r Line30

Liquid Line20

10

00 10 20 50 90 10030 40 60 70 30

Time (min)


Analysis of Experimental Data

LHP start-up represents one of the most complex transient phenomenons in LHP

operation. Start-up is a function of evaporator and compensation chamber construction,



and is not only dependent on fluid inventory but also on initial conditions inside the

evaporator core and wick, as well as operation immediately prior to start-up. To that end,

a strong effort was made to keep the initial states of the fluid within the loop as similar as

possible between tests by following the same procedure and setup, as described in

Section 5.3. It is however difficult to completely control the states of the fluid in each

section of the LHP prior to start-up, and is taken into account when examining the data.

The results of the first two fluid charges show a high degree of instability within the

loop during start-up as vapour begins to flow through the vapour line. Figure 5-15 and

Figure 5-16 both indicate that shortly after vapour begins to flow through the vapour line,

the evaporator experiences a sudden cooling effect of 2 to 3 °C from cool liquid arriving

from the liquid line. At the same time, the liquid line experiences a similar cooling effect

due to liquid being pushed from the condenser. However, the cooling effect is not

sustained and temperatures of all components then begin to rise. Steady-state

temperatures were only achieved by re-applying power after allowing the system to cool

below a temperature of 40 °C. It is possible that such characteristics may be a result of a

large temperature overshoot caused by vapour phase in the evaporator core, liquid phase

in the vapour line, or both. If the vapour line is flooded with liquid, a relatively large

superheat is required to initiate nucleate boiling and displace the liquid in the grooves. If

on the other hand, there is a significant amount of vapour in the evaporator core, the

temperature of the evaporator and compensation chamber will rise until cold liquid from

the condenser can compensate for the increased heat leak. Another possible explanation



is that vapour blockage in the pores of the wick may be limiting the available area of

vaporization. This would not only increase the effective thermal resistance of the system

but also reduce the flow of vapour in the vapour line and mass flow rate of fluid in the

system. A reduced mass flow rate would result in less subcooling, which is necessary for

offsetting the heat leak from the evaporator to the compensation chamber. This in turn

would increase the LHP operation temperature beyond the maximum allowable

temperature. The results of the last two fluid charges (20 g, 28 g) produced more reliable

and consistent results in comparison to the first two fluid charges (15 g, 23 g). A fluid

charge of 28 g indicated some degree of vapour blockage in the wick since the

temperature of the evaporator was relatively high in comparison to the temperature of the

vapour line and compensation chamber over the entire range of applied power. However,

there was still sufficient fluid flow in the loop to provide the necessary amount of

subcooling to achieve steady-state temperatures. A fluid charge of 20 g appeared to

produce the most desirable results. No significant temperature overshoot was observed

and generally, steady-state temperatures immediately after start-up appeared to be lower

compared to the other fluid charges. The only significant difference found between each

fluid charge was the time required for initiation of circulation of the fluid. For example,

it took approximately 10 minutes from the time power was turned on (15 W) to

circulation within the loop for a fluid charge of 20 g. In comparison, circulation of fluid

in the loop was initiated 22 minutes after applying the same power for a fluid charge of

28 g. In general, it was found that start-up times were shorter for smaller fluid



inventories. A summary of start-up times for three of the four fluid charges is provided

in Table 5-3. It is suggested that start-up times are reduced because a smaller mass flow

rate is required to displace liquid within the lines since there is less liquid to displace for

smaller fluid inventories.

Table 5-3: Summary of Start-up Times

Applied P ow er (W)15 g, Start-up

Tim e (hr:min:s)2 0 g, Start-up

Tim e (hr:min:s)23 g, Start-up

Tim e (hr:min:s)28 g, Start-up

Tim e (hr:min:s)

5 0:17:36 0:24:36 0:25:17 2:20:27

10 0:15:17 0:19:11 0:24:24 0:52:00

15 0:07:48 0:10:06 0:02:12 0:22:49

20 0:06:02 0:04:54 0:09:00 0:10:28

In general, the effect of fluid inventory does not appear to be a dominant factor for

start-up. The state of the working fluid within the system and wick just prior to start-up

appears to have a much greater impact. However, it is noted that the amount of working

fluid can impact the amount of superheating or time required before start-up, since less

fluid will typically result in less liquid collection within the vapour grooves at any given

time. As will be discussed in the following section, fluid inventory has a much greater

impact on the steady-state operating temperature and maximum heat transfer of the LHP.

5.5.3 Steady-State Operating Temperatures

After LHP startup was observed, power cycling tests were employed to study the effect

of fluid inventory on LHP operational stability, by handling changes in heat loads. Tests

involved successive application of increasing heat loads while maintaining a constant

sink temperature of 5 °C. As discussed in the operating procedures, temperatures were



allowed to reach steady state before each increase in power. Steady-state was defined as

a temperature change of no more than 0.5 °C over a period of at least 20 minutes. Figure

5-20 and Figure 5-21 illustrate the temperature profiles obtained during power cycling

tests for fluid charges of 20 g and 23 g respectively. For each fluid charge, stable

operation was obtained between 10 W and 50 W without de-priming or significant

temperature overshoot. Both curves indicate that the operating temperature increases

with input power in an approximate linear relationship, suggesting that the LHP was

operating in constant conductance mode for both cases. As will be shown, the other two

fluid charges presented similar results, but in a much narrower power range.

Evapora tor V a p o u r Line


50 0-

P o w e r - 3 0

8 0 100 Time (min)

120 140 160 180

Figure 5-20: Temperature Profile during Power Cycling Tests for 20 g Fluid Inventory



Evapora to r V a p o u r Lino


P o w e r

100 150 200 25 0 T im e (min)

300 3 5 0 400

Figure 5-21: Temperature Profile during Power Cycling Tests for 23 g Fluid Inventory

Additional power cycling tests were used to generate performance curves for each fluid

charge. The resulting performance curve for each fluid charge is shown in Figure 5-22.

The results represent the average steady-state temperature of the evaporator at each

power load. Figure 5-22 illustrates that LHP operation is very sensitive to fluid charge.

It was observed that steady-state operating temperature and maximum/minimum heat

transfer capacity of the LHP varied significantly between the highest and lowest charges

(15 g, 28 g) and the nominal charges (20 g, 23 g). The LHP demonstrated successful

operation between an applied power of 10 W and 50 W for fluid charges of 20 g and 23

g. In comparison, the LHP demonstrated successful operation between a much narrower

power range of 5 W and 20 W for fluid charges of 15 g and 28 g.



80 75 70

O 650£ 601 55 0)I 50 ,2 45

40 35 30

0 10 20 30 40 50 60Power (W)

■ 15 g♦ 2 0 g a 23 g• 28 g

Figure 5-22: Performance Curve for each Fluid Inventory

The specified power range indicates the maximum and minimum heat transfer capacity of

the LHP for each fluid charge. The heat transfer capacity was considered to reach a

maximum if the temperature at the evaporator surface began to exceed the maximum

allowable temperature of the wick (>75 °C). The minimum power required to sustain

stable operation yielded the minimum heat transfer capacity of the LHP. An interesting

feature of Figure 5-22 is that the smallest fluid charge of 15 g results in the lowest steady-

state operating temperature at low heat loads. Specifically, the LHP operated almost 10

°C cooler compared to a fluid charge of 28 g for a power input of 5 W. Furthermore, it

was the only fluid charge to successfully start-up and sustain loop operation at 2 W.

However, tests also revealed that the maximum heat transfer for this particular charge

was limited to 20 W. Overall, a lower fluid charge of 15 g results in a very steep



performance curve for this particular design, where start-up is easily initiated at low

powers because of less liquid to displace, and is also susceptible to early dry-out at high

powers due to insufficient subcooling. Similar results were obtained for a fluid charge of

28 g. In this case, however, a relatively low maximum heat transfer is not the result of

wick dry-out. Instead, a large fluid inventory may result in hard-filling the reservoir. In

general, the vapour-liquid interface moves towards the end of the condenser to provide

the required subcooling at high heat loads. For this particular design, a subcooler was

added to the liquid line since the available area for heat transfer in the condenser was

relatively small. Therefore, at high heat loads, the vapour-liquid interface may move into

the liquid line to provide sufficient cooling. When this happens, the liquid from the

condenser and part of the liquid line is displaced into the compensation chamber. If the

vapour-liquid interface moves too far into the liquid line, the compensation chamber will

become hard-filled and block the flow of cool liquid. Consequently, the liquid inside the

compensation chamber will stagnate and increase in temperature due to parasitic heating.

Heat leak from the evaporator to the compensation chamber will then dominate, causing

the operating temperature of the LHP to increase. The most effective fluid charge for this

particular design was found to be 20 g. With a fluid inventory of 20 g, the LHP

demonstrated successful operation between a power range of 5 and 50 W. Even at a

power load of 50 W, the resulting operating temperature of 58 °C was well below the

critical cut-off temperature, indicating that the LHP was capable of handling higher heat

loads. Similar results were obtained for a fluid charge of 23 g, however, the performance



curve increased at a much faster rate for applied heat loads higher than 20 W. At an

applied load of 50 W, the temperature difference between the two curves was almost 15

°C. The difference in temperatures may lie in the difference between start-up

characteristics for each charge. If vapour bubbles are still present in the wick after start

up, then the operating temperature of the LHP may increase for the same applied power.

For a fluid charge of 23 g, it is also possible that similar conditions to those found for a

fluid charge of 28 g may be present in the compensation chamber at high heat loads.

5.5.4 Effective Thermal Resistance

LHP efficiency can also be examined in terms of effective thermal resistance. The

effective thermal resistance of an LHP is a function of the steady-state temperature of the

evaporator and condenser and is defined as:

R f f = = -£----------------------------------------------(5.1)Q Q ,

Figure 5-23 demonstrates the effective thermal resistance of the LHP for each fluid

inventory, based on the results obtained from power cycling. It should be noted that the

average temperature of the evaporator and condenser was used for calculating the thermal

resistance. The evaporator thermocouple was situated at the centre of the evaporator

while the condenser thermocouple was situated near the condenser exit. In general, this

assumption is more accurate when the condenser is completely utilized and the average

temperature approaches the liquid temperature at the condenser exit (Maidanik et al.,

2005).



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a)>oa>LU

14

12

10

8

6

4

2

0

■

__ • ___♦■

--------- ---------- ----------

i ■ •#

* -

1

11

1 * 41 A i► A (►

♦ 20 g

10 20 30 40Power (W)

50 60

Figure 5-23: Effective Thermal Resistance for each Fluid Inventory

The data indicates that the effective thermal resistance decreases as the heat load to

the evaporator increases. As expected, the minimum thermal resistance observed during

testing occurred for a fluid charge of 20 g and was approximately 0.951 °C/W.

Variations in LHP thermal resistance between each fluid charge were a result of

differences in the evaporator thermal resistance. The thermal resistance of a LHP can

also be defined in terms of the heat transfer coefficient and the active surface area as

follows:

= —*— + ----"a , S ., a S.

(5.2)c cond

where:

ot„ is the heat transfer coefficient of the evaporator



acond is the heat transfer coefficient of the condenser

Se is the active area of the evaporator

SCond is the active surface of the condenser

The first component determines the thermal resistance of the evaporator while the second

component determines the thermal resistance of the condenser. In general, the active

surface area of the condenser and heat transfer during condensation is quite large in

comparison to the evaporator. As such, the condenser term in Eq. (5.2) has minimal

impact on the effective thermal resistance of a LHP. It therefore becomes clear that in

order to decrease the effective thermal resistance, the heat exchange in the evaporator

must be maximized. That means that in cases where there is wick dry-out or the

compensation chamber is hard-filled, the total thermal resistance of the system will

increase. This is evident at high heat loads for fluid charges of 15 and 28 g.

5.5.5 Heat Transfer Coefficient

The heat transfer coefficient in the evaporator is another useful characteristic to help

quantify efficiency of a LHP. The heat transfer coefficient is calculated as follows:

a - - A (5 3 )

where ATe = Te ~TVii is the temperature difference between the evaporator and the

vapour line inlet, and Se is the active evaporator area. Ideally, the temperature of the

evaporator should be measured at the inner wall surface of the casing. However, since it

is difficult to position thermocouples on the inner surface of the evaporator, the external



wall temperature was used instead. Figure 5-24 illustrates the heat transfer coefficient for

each fluid charge. As expected, the heat transfer coefficient increases with increasing

power, based on Eq. (5.3). An interesting feature is that the heat transfer coefficient

experiences an abrupt change as the LHP approaches its critical operating temperature.

For example, the heat transfer coefficient is reduced from 261 to 248 W/m2K as the

applied power is increased from 45 to 50 W for a fluid charge of 23 g. Similar results

were observed for fluid charges of 15 and 28 g. These results suggest that the evaporator

experienced a sudden increase in thermal resistance for three of the four charges as each

approached the maximum cut-off temperature. This also suggests that the LHP was not

tested to its full potential with a fluid charge of 20 g.

_ 350*

N

E 300 §*T 250 .2 £ 200 <D0 150 0)£ 100 2

50re a>1 0

0 10 20 30 40 50 60Power (W)

i : < ►

♦;__a

---------- i

------f----*iiiii

T t> A :

i

* .!

■

i "

iiiiiiii ---------

■ 15 g♦ 20 g A 23 g• 28 g

Figure 5-24: Heat Transfer Coefficient for each Fluid Inventory



5.6 Test Phase 3: Experimental Study of LHP Operating Characteristics

For the current LHP design, a fluid charge of 20 g produced the best all around and

consistent results during the second phase of testing. As a result, a fluid charge of 20 g

was selected for the final phase of testing. The primary objective of the final phase of

testing was to study the effect of elevation and periodic heating on LHP performance.

Additionally, tests were conducted to investigate the possibility of temperature hysteresis.

5.6.1 Effect of Elevation

The effect of elevation on LHP performance was studied for four configurations: 5°

positive elevation, no elevation, 5° degree adverse elevation, 10° adverse elevation.

Elevation refers to the position of the evaporator with respect to the condenser. An

adverse elevation means the evaporator is positioned above the condenser while a

positive elevation means the evaporator is positioned below the condenser. The steady-

state operating temperature profiles for all four elevations are shown in Figure 5-25.



70

60

£ 50 a> 40 -*-»

(5 30 aa 20 H

10

0

■ Zero Elevation ♦ 5° Positive Elevationa 5° Negative Elevation *10° Negative Elevation

----------- ..... ------------------------------------------- -------I I I I

_______ # - i .| A ♦ * ♦• i ♦ *

I l l * ’♦ ^ 1 1 1 11 1 1 I1 1 1 1

> 1 1 1 I I I ! 1 1 1 11 1 1 1 1 1 I 11 1 1 11 1 1 1 1 1 1 I

...... ....... 1............. 1-----------1-----------1------- ---10 20 30

Power (W)

40 50 60

Figure 5-25: Effect of Elevation on Steady-State Operating Temperatures

The overall trend of the data suggests that elevation has a significant influence on

performance at low powers. Figure 5-25 shows that steady-state temperatures increased

with increasing adverse elevation, up to an applied power of 20 W. The increase in

temperature can be explained as follows. As the pressure difference across the wick

increases due to gravity head, the difference in saturation temperatures must also

increase, as required by Eq. (2.7). Since the liquid enthalpy entering the compensation

chamber does not change, an increase in evaporator vapour temperature must occur to

satisfy the increasing pressure drop. An increase in evaporator temperature leads to a

higher heat leak which in turn increases the operating temperature of the LHP.

Therefore, higher adverse elevations result in an increase in heat leak which must be

compensated for by an increase in steady-state operating temperature for low heat loads.



At higher heat loads, heat leak is compensated for by an increase in liquid subcooling,

m CpA T , through an increase in mass flow rate. Thus, at higher heat loads, gravitational

pressure losses have an almost negligible effect on steady-state operating temperatures.

This was demonstrated for zero elevation and 5° positive elevation. Both curves begin to

converge immediately and follow approximately the same trend for heat loads above 30

W. It is noted that the temperature profiles for 5° and 10° adverse elevation also

converge immediately. However at approximately 20 W, they begin to diverge from the

other two curves and rise much faster in temperature. The rise in temperature may be due

to partial dry-out of the wick since the current design does not include a secondary wick

or bayonet tube to directly supply fluid to the primary wick.

5.6.2 Periodic Heating

The main purpose of periodic heating was to investigate the effect of rapidly varying heat

loads, thus simulating conditions which may be encountered during operation such as a

spacecraft entering into eclipse or variable power dissipations from a payload or

electronic devices. For these tests, power was increased from 0 W to 15 W in constant

intervals of 10 minutes after startup. As demonstrated in Figure 5-26, rapid power

changes did not have any negative effect on the performance of the LHP. In fact, the

LHP experienced a temperature overshoot of almost 10 °C at start-up compared to the

average temperature observed during periodic heating. This suggests that the working



fluid may have been in an unfavourable state within the loop prior to startup, as described

in Section 5.5.2.

E v ap o ra to r V ap o u r Line

Liquid LineC o m p e n sa tio n C h a m b e r

100 150 T im e (min)

200 250 300

Figure 5-26: LHP Operation During Periodic Heating. 15 W, 10 min cycle

5.6.3 Temperature Hysteresis

Power cycling tests were also used to investigate whether or not the current design

displays characteristics typical of temperature hysteresis. Temperature hysteresis is

identified when the loop operating temperature depends on the recent history of the

applied power even when all other test conditions, including ambient and sink

temperature, remain constant. As discussed in Section 2.2.3, there is strong evidence that

temperature hysteresis is directly related to the void fraction of the evaporator core (Kaya

et al., 1999). An accumulation of vapour bubbles in the evaporator core increases heat



leak between the evaporator and compensation chamber thus resulting in a higher

operating temperature. It has also been reported in literature that in designs without a

secondary wick, temperature hysteresis may be caused by tiny vapour bubbles entrapped

in the wick that cannot completely collapse or vent to the compensation chamber (Ku,

1999). Experimental data obtained during power cycling supports this hypothesis. As

shown in Figure 5-27, temperature hysteresis in the current design was evident. In this

example, the input power to the evaporator was cycled in steps of 10 W.

E v ap o ra to r V a p o u r Lino

Liquid LineC o m p e n sa tio n C h am b er

o 50

4 0 5

CL

P o w e r

100 150 200 250Tim e (min)

Figure 5-27: Temperature Hysteresis

Additional tests for hysteresis produced similar results, especially after large steps in

power (> 10W). These results suggest that for the current design, a significant amount of

vapour bubbles collect in the polyethylene wick after large steps in applied power. Such



observations have also been made when testing CPLs with polyethylene wicks (Nikitkin

and Cullimore, 1998). Vapour bubbles in the wick can block liquid return paths,

resulting in partial dry-out of the wick thus reducing the effective area of vaporization

and increasing the thermal resistance of the evaporator. This leads to sharp increases in

loop operating temperature and in some cases loop de-priming. These results were

similar to those obtained during start-up tests, described in Section 5.5.2. In many cases,

it was found that the LHP required some ‘pre-conditioning’ to remove any vapour

bubbles from the wick. Specifically, an excessive heat load was applied to the evaporator

to collapse the vapour bubbles in the wick. Once the loop reached its maximum cut-off

temperature, the heat load was removed until the temperatures dropped below a certain

value. This method proved to be successful since steady-state temperatures were

obtained.


Chapter 6: Mathematical Modeling

6.1 Software Development

As discussed in Chapter 3, the software package EASY 2000, developed by TAIS Ltd.,

was used to help optimize initial LHP designs. However, due to its limitations, in-house

software was required and was subsequently developed in parallel with the final LHP.

Developing in-house software avoids the use of some black-box solutions introduced

with EASY 2000 and provides greater flexibility during design and reduces long term

costs. Due to time constraints, efforts were focused on developing the software necessary

for completing the LHP design and testing portions of this study. Numerical models

were developed in Matlab and provide functions such as estimating properties of various

working fluids, calculating fluid inventory, and sizing the compensation chamber. These

models provide the foundation for future work on 1-D steady-state performance solvers

(see Appendix A for details). Details of each model are provided below.

6.1.1 Graphical User Interface (GUI) & Input/Output Parameters

The Matlab Graphical User Interface Development Environment (GUIDE) was used to

develop a GUI to simplify the process of entering input data, calculating LHP

characteristics, and extracting the output results. The software is organized into a series

126


CHAPTER 6: MATHEMATICAL MODELING 127

of modules, with each module providing a specific function such as calculating properties

of a specific fluid or calculating and displaying the figure of merit. As shown in Figure

6-1, the main toolbar provides an interface for the user to access each individual module.

[fflthpPw fr ** ‘ « - «. »—j gFife Edit O ptions f f f j i j g Help ~~ ~ ~ ’ ” " ” ' ~

j Fluid Properties

j Fluid Inven to ry

Merit Number

{ LHP P ressu re Drop

I LHP Perform ance

Sensitivity Studies

Figure 6-1: LHP Design Software Toolbar

Typically, the first task is to enter the physical properties and conditions of the LHP.

This is accomplished by accessing the edit properties form from the ‘Edit’ menu on the

main toolbar. As illustrated in Figure 6-2, input parameters include the detailed

dimensions of each component, wick porosity, wick permeability and wick effective pore

radius. It also includes external conditions such as the sink temperature, cooling water

volume flow rate, ambient temperature, and external thermal conductance of the

condenser. It is noted that the external thermal conductance of the condenser refers to the

thermal conductance per unit length from the inner surface of the condenser tube to the

outer surface of the condenser plate and can typically vary between 1 and 20 W/m-K,

depending on the flow arrangement and thermal contact resistance between the condenser

tube and the condenser plate. All input data is saved as a .txt file under a specified

directory (Matlab working directory by default) and is editable.



M 1 i l i r P io p n h i”

j File

—Wick Properties —

Wick Density [kg/m3]: j 965

Pore Radius [m]: f~ 0 00002

Wick OD [m]: } 0.022225

Wick ID [m). j 0 . o i 5 ~

Porosity [Q..1]: f - 03

Wick Bore Length [m]: | 0.1134125

Wick Un-Bored Length fm]. | 0.0015875

Secondary Wick Length [m]: \ 0

Secondary Wick ID [m\. j ET~

- Evaporator & Groove Properties -

Heated Length [m]: 0.03

Wall Thickness [m]: | 0.0015875

Number of Axial Grooves: f 4

Groove Depth [m]: j 0.0015

Groove Width [mj. \ O. OOG

Number of Circumferential Grooves:

Groove Depth fm]: \ 0.0015

Groove Width [m]: 1 0 001778

Land Width [m]: 0 022222

— Vapour Line Properties---------------------------- — Condenser Properties---------------------------- — Liquid Line Properties-----------------------------

Lice Length lm]: f 6.381 Line Length [m]: f ^-381 Line Length [m]: \ 0.381

''1Hydraulic Diameter [ml | 0 004572 Hydraulic Diameter [ml: I 0.004572 Hydraulic Diameter H : ( 0.004572

WallThickness [m]: | 0.000889 WallThickness [m], ) 6.000888 WallThickness [ml | 0.600888

Figure 6-2: Edit Properties

As shown in Figure 6-1, each calculation module is available under the ‘Tools’ menu and

currently includes algorithms to calculate fluid properties, fluid inventory, and merit

number. Additionally, the user can study the effects of varying the input parameters.

The program is also designed for easy integration of future modules.



6.1.2 Fluid Properties

The fluid properties module allows the user to review relevant working fluid properties

for a specified saturation temperature. As shown in Figure 6-3, the trend of each property

over a range of saturation temperatures is also displayed.

R S W o i H i w I l u i d l ' i o p e i l i e s !'Q -Close

Working Fluid: J Acetone

Working Temperature Flange (min/max) [Degrees Celcius]:

Temperature R ange of Diagram (min/max) [Degrees Celcius]:

-40

-40

140

] 140Ts= | 2Q Degrees Celcius

C Saturation Pressure, [Pa] | 23236

x 10

r Liquid Density. [kg/m3] ) 787 847

C Vapour Density, [kg/m3] | 0 61553

r Liquid Viscosity, [m2/s2] J OJ066322475

<• Vapour Viscosity [m2/s2] | & 81^5e-006

C Surface Tension [N/m] | 0 0227757

C Latent Heat [J/kg] j

C Liquid Conductivity [W/mK]

530450

0.172103100 150

T e m p e ra tu re [C e lc iu s ]

Calculate

Figure 6-3: Working Fluid Properties Calculator

The relevant properties of the working fluid are saturation pressure, liquid density,

vapour density, liquid viscosity, vapour viscosity, surface tension, latent heat, and liquid

thermal conductivity. The algorithms used in this module to calculate each working fluid

property are also accessed by other modules to accurately calculate fluid inventory and



compensation chamber volume. Each fluid property was expressed as a function of the

LHP operating temperature (saturation temperature, Tsat) and was curve fitted into a fifth-

order polynomial with errors of approximately 1-5%.

The coefficients of the polynomials were obtained from (Faghri, 1995). The model

currently includes coefficients for four working fluids: acetone, ammonia, methanol, and

water. Coefficients for other fluids can easily be added to the module.

6.1.3 Figure of Merit

As discussed in Section 2.3.1, the liquid density, surface tension, liquid thermal

conductivity, and liquid viscosity of a working fluid are combined to form a

dimensionless parameter known as the merit number or liquid transport factor. The merit

number can be used evaluate and compare the effectiveness of various working fluids at

specific operating temperatures. It is helpful in selecting the appropriate working fluid

when designing a LHP for a specific temperature range and task. The merit number was

defined as:

(6 .1)n=1

where:

Z is the fluid property

is the corresponding coefficient

(6 .2)



The merit number module produces a figure of merit as shown in Figure 6-4. The user

can select which working fluids to plot and the temperature range.

Hl _ _ .

: !: s A , . v . y : ; K W ® ' . . . . . : " / ; V M

Figure of Merit for Various Working FluidsS

5

AcetoneAmmonia

Methanol W ater

4—O3

Jj

-*100 0 50 100 150 200 250-50

Temperature [°C]

Figure 6-4: Figure of Merit

6.1.4 Fluid Inventory

The fluid inventory module was designed to estimate the fluid charge and compensation

chamber volume. The module includes code for both calculation methods described in

Section 2.3.4 and allows the user to select between the two. The user can specify the

working fluid, operating temperature range, and compensation chamber void volume for

each method. By default, the code is setup to concurrently estimate the fluid charge and



compensation chamber volume. However, it is possible to calculate the fluid charge

separately by specifying the dimensions or volume of the compensation chamber. Figure

6-5 demonstrates an example of the output produced by the fluid inventory module.

RJChargeSolveiS :r: lj; .'Mi; "s'

i | A cetone "■, ] K = j 1.2 I min ["C]: f~ 5

Tmax{°C]: | 70

Diameter of Reservoir [m] = j 0.0211836

Reservoir Length [m] = 0.0821916

Volume of Reservoir [cc] = 43.3346

Volume of LHP = 79.9249

Volume of Vapour Line = 15.0664

Volume of Liquid Line = 6 255

Volume of C ondenser = 6.255

Volume of Pores in W ick = 4 .43529

Volume of Axial G rooves = 4.14

Volume of Circumferential G rooves = 0.438556

M ass of Fluid C harged [g] =

Note: All volumes a re in c c } Caicuiaije j

Figure 6-5: Fluid Inventory and Compensation Chamber Sizing


Chapter 7: Conclusions and Recommendations

7.1 Summary and Conclusions

This thesis presented a theoretical and experimental analysis of LHPs. In particular, it

discussed the procedures of designing and developing LHPs for testing. Additionally, an

experimental investigation of LHP performance was also discussed. The purpose was to

establish a solid understanding of LHP technology and to demonstrate its flexibility,

reliability, and robustness for a wide range of thermal control applications.

In this study, three LHPs were designed, manufactured and tested in sequence, and

with each new design incorporated design changes based on test results obtained from the

previous unit. Initially, components were selected based on a set of criteria which

included flight heritage, cost, availability, compatibility, and performance characteristics.

Each component was then optimized to minimize pressure losses and therefore maximize

the potential heat load limit. Each unit was subsequently constructed following a strict

procedure which included machining, cleaning, assembly, evacuation, charging, and

sealing.

An experimental investigation of LHP performance characteristics was performed in

three phases: investigate LHP functionality; investigate sensitivity of LHP performance

with respect to fluid charge; and study the effect of elevation and periodic heating on the

133


CHAPTER 7: CONCLUSIONS AND RECOMMENDATIONS 134

maximum transport capacity and stability of the LHP. In the first phase of testing, the

first test unit demonstrated clear start-up characteristics and stable operation. However,

over time the performance of the unit degraded until it was no longer functional.

Inspection of the wick revealed minor damage to the outer layer of the material, possibly

due to operation beyond its maximum operating temperature. After reviewing the results,

the second LHP was designed with minor modifications to increase performance.

Unfortunately, test results of the second unit revealed no significant fluid flow in the

loop. Upon examination, it was found that the LHP failed due to small leaks at the

Swagelok interfaces. With consideration of these findings, the final LHP was constructed

using seamless stainless steel tubing with much greater tolerances. Furthermore, a

subcooler was added to the liquid line to increase liquid subcooling in the compensation

chamber to enable testing in a wider range of powers. Preliminary tests demonstrated

clear start-up and excellent performance. Continuous and stable operation was also

demonstrated. The third test unit was therefore selected for the second phase of testing.

In the second phase of testing, sensitivity studies were performed to evaluate the

effect of fluid inventory on LHP performance. In particular, an effort was made to

demonstrate the effect of fluid inventory on start-up transients, steady-state operating

temperature, maximum transport capacity, effective thermal resistance, and heat transfer

coefficient. A total of four fluid charges were selected for analysis. These fluid charges

each corresponded to a specific void fraction in the compensation chamber during worst

case hot conditions and a specific liquid-filled fraction in the compensation chamber



during worst case cold conditions. Power cycling tests revealed that fluid inventory is

not a dominant factor in start-up but may impact the time required for initiation of fluid

circulation within the loop. Specifically, smaller fluid charges were found to require less

time for initiation of fluid flow since less liquid was available to collect in the vapour

grooves. Alternatively, higher fluid charges took significantly more time for initiation of

fluid flow, possibly due to excess liquid collection in the vapour grooves which increase

the amount of superheat required. Additionally, fluid inventory was found to have a

significant impact on steady-state temperatures and maximum/minimum heat transfer

capacities. Specifically, the lowest fluid charge of 15 g resulted in the lowest steady-state

operating temperatures, but only for small power loads. As the input power was

increased, the steady-state temperature of the loop rose sharply due insufficient liquid

supply for subcooling. Consequently, wick dry-out and loop deprime was evident at

power loads greater than 20 W. Similar results were demonstrated for the highest fluid

charge of 28 g. In this case, LHP performance was also weak at higher powers and was

most likely the result of hard-filling the compensation chamber wherein the liquid-vapour

interface moves towards the condenser exit and liquid is pushed into the compensation

chamber. If the fluid charge is too high, there may not be sufficient room in the

compensation chamber to compensate for the excess liquid. In such a case, the liquid

stagnates and increases in temperature due to parasitic heating. This in turn increases the

loop operating temperature. It was determined that for this specific design, a fluid

inventory of 20 to 23 g produces optimum results. In particular, a fluid charge of 20 g



produces excellent performance characteristics and was therefore selected for the final

phase of testing.

The final phase of testing was performed to study the effect of elevation and periodic

heating on LHP performance, as well as test for the presence of temperature hysteresis.

Initial tests revealed that elevation can have a significant impact on loop performance at

low powers. Specifically, the steady-state temperature of the loop was found to increase

with increasing adverse elevation for the same power load, up to 20 W.

This effect decreased with increasing power, as each performance curve converged. It is

noted, however, that at higher elevations the LHP had an increased tendency to deprime

for power loads greater than 20 W. This was attributed to the fact that there was no

secondary wick or bayonet tube to help supply the primary wick with fluid. Periodic

heating tests were used to investigate the effect of rapidly varying heat loads, thus

simulating conditions which may be encountered during typical operation. It was found

that rapid changes in power did not have any negative effect on LHP performance. In

fact, in many cases it helped free trapped vapour bubbles in the wick thus decreasing the

average temperature of the evaporator. Finally, power cycling tests were performed to

investigate temperature hysteresis. Temperature hysteresis was identified when the loop

operating temperature varied for the same applied power, even when all test conditions

including ambient and sink temperature remained constant. Several instances of

temperature hysteresis were observed during testing. It was concluded that such

characteristics may be a result of the void fraction in the evaporator core or excess vapour



bubbles trapped in the wick structure. Vapour bubbles in the wick may block liquid

return paths, resulting in partial dry-out of the wick and an increase in temperature.

Overall, the results of the study suggest that the techniques and procedures used for

designing and manufacturing a LHP were successful. After optimizing the fluid

inventory, the LHP demonstrated reliable start-up and stable operation between a power

load range of 5 to 55 W. Additionally, the LHP showed no adverse effects due to

periodic heating. There were, however, some inconsistencies during operation.

Specificlaly, temperature hysteresis was sometimes visible during power cycling tests.

This issue may be resolved by using a more efficient primary wick or by including a

secondary wick. These results provide a solid base for future work towards developing

smaller, more efficient LHPs for spacecraft thermal control and electronics cooling.

7.2 Recommendations

Based on the results of this study, the following recommendations were made for future

work.

L H P Laboratory a t Carleton U niversity

The LHP laboratory at Carleton University is adequately equipped for manufacturing and

testing LHPs. However, there are several options which may be explored to improve the

facility and increase the efficiency of testing. They are as follows:

• Design and build a new test frame that can accommodate a wider range of LHP

designs and can more accurately determine LHP orientation;



• Improve data acquisition by automating test processes such as varying heater

power and automatic shut-down; and

• Design and construct an environmental chamber with the capability to control

ambient conditions. It is suggested that the chamber be designed to control the

ambient temperature and simulate both forced and natural convection conditions.

L H P D esign

The following highlights some key areas of interest regarding loop performance or data

extraction and are based on the results obtained from this study:

• Design and construct a more efficient condenser and heat exchanger, and

increase the active cooling area of the condenser;

• Construct a new evaporator saddle that can accommodate a second heater strip

and more effectively distribute the applied power, which is currently limited to

60 W;

• Develop an LHP with instruments to measure pressure and mass flow rate within

the loop. As well as, increase the number of temperature measurements along the

loop; and

• Begin development of smaller designs based on the results of this study, starting

with a V2” evaporator. In addition, evaluate the feasibility of employing current

manufacturing and assembly techniques.



P erform ance T esting

Detailed testing may be performed in a number of new areas, including:

• Active temperature control of the compensation chamber using a thermal electric

cooler (TEC);

• Study the effect of varying ambient and sink temperatures;

• Study the effect of positive elevation and tilt;

• Study the effect of wick structures with varying effective thermal conductivities;

and

• Conduct a more in depth examination of start-up transients.

M athem atica l M odeling

The models developed for estimating properties of various working fluids, calculating

fluid inventory, and sizing the compensation chamber were adequate for this study.

However, future development of an algorithm to model the steady-state performance of a

LHP in variations conditions and orientations is necessary. The following is a list of

recommendations to achieve this goal:

• Develop and incorporate a LHP steady-state performance model in the current

Matlab models, and make accessible through the GUI;

• Perform experimental validation of the model (ie., compare results of the model

to those obtained through experimental testing); and



• Perform parametric study of LHP performance characteristics using the steady-

state model. Areas of interest may include effect of sink temperature, ambient

temperature, elevation, heat leak, and condenser conductance.


References

Chi, S. W. (1976). Heat pipe theory and practice : a sourcebook. Washington, Hemisphere Pub. Corp.

Chuang, P. Y. (2003). An Improved Steady-State Model o f Loop Heat Pipes Based on Experimental and Theoretical Analyses. Department of Mechanical and Nuclear Engineering, Pennsylvania State University. Doctor of Philosophy: 271.

Dunn, P. D. and D. A. Reay (1982). Heat pipes. Oxford ; New York, Pergamon Press.

Faghri, A. (1995). Heat pipe science and technology. Washington, DC, Taylor & Francis.

Holman, J. P. (1990). Heat transfer. New York, McGraw-Hill.

Kaya, T. and T. Hoang (1999). Mathematical Modeling of Loop Heat Pipes and Experimental Validation. Journal o f Thermophysics and Heat Transfer 13(3): 314-320.

Kaya, T. and J. Ku (2003). Thermal Operational Characteristics of a Small-Loop Heat Pipe. Journal o f Thermophysics and Heat Transfer 17(4): 464-470.

Kaya, T., J. Ku, T. Hoang, K. Cheung (1999). Investigation of Low Power Start-Up Characteristics of a Loop Heat Pipe. Space Technology and Applications International Forum.

Khrustalev, D. (2001). Loop Heat Pipe Technology for Electronics Cooling. International Conference on High-Density Interconnect and Systems Packaging.

Kim, J. and E. Golliher (2002). Steady State Model of a Micro Loop Heat Pipe. 18th IEEE Semi-Therm Symposium, JINHA Sciences, Inc.

Ku, J. (1993). Overview of Capillary Pumped Loop Technology. ASME 29th National Heat Transfer Conference, Atlanta, GA,.

Ku, J. (1999). Operating Characteristics of Loop Heat Pipes. 29th International Conference On Environmental Systems, Denver, USA, Society of Automotive Engineers.

141


Lashley, C., S. Krein, P. Barcomb (1998). Deployable Radiators - A Multi-Discipline Approach. Society o f Automotive Engineers 981691.

MacDonald, E. (2004). Experimental and Numerical Investigation o f Passive Two-Phase Heat Transfer Devices for Space Applications. Mechanical and Aerospace Engineering. Ottawa, ON, Carleton University. Master of Applied Science: 133.

Maidanik, Y. F. (2003). Loop Heat Pipes - Development and Application. 7th International Heat Pipe Symposium, Jeju, Korea.

Maidanik, Y. F. (2005). Review - Loop Heat Pipes. Applied Thermal Engineering 25: 635-657.

Maidanik, Y. F., S. Vershinin, J. Dolggirev (1985). Heat Transfer Apparatus. United States. 4515209.

Maidanik, Y. F., S. V. Vershinin, M. A. Korukov, J. M. Ochterbeck (2005). Miniature Loop Heat Pipes - A Promising Means for Cooling Electronics. IEEE Transactions on Components and Packaging Technologies 28(2): 290-296.

Nikitkin, N. and B. Cullimore (1998). CPL and LHP Technologies: What are the Differences, What are the Similarities? 28th International Conference on Environmental Systems, Denver, USA, Society of Automotive Engineers.

Pastukhov, V. G., Y. F. Maidanik, C. V. Vershinin, M. A. Korkov (2003). "Miniature Loop Heat Pipes for Electronics Cooling." Applied Thermal Engineering 23: 1125-1135.

Peterson, G. P. (1994). An introduction to heat pipes : modeling, testing, and applications. New York, Wiley.

Riehl, R. and T. Dutra (2005). Development of an Experimental Loop Heat Pipe for Application in Future Space Missions. Applied Thermal Engineering 25: 101-112.

TAIS (2000). EASY2000. Moscow: Application Program Package Manual.


Appendix A

A .l 1-D Steady-State Analytical Model

Future work on a steady-state performance solver will be required to help improve the

efficiency of current LHP designs. A short study of the mathematical models formulated

by Kaya and Hoang (1999) is provided below and is intended as a starting point for future

work. The model is ideal for estimating performance characteristics such as the pressure

and temperature distributions along the LHP, mass flow rate, heat leak, and liquid

subcooling as a function of input power for given LHP conditions such as sink

temperature and ambient temperature. The algorithm has proven success ((Kaya and

Hoang, 1999), (Chuang, 2003), (Riehl and Dutra, 2005)), and has high flexibility with the

ability to incorporate more accurate two-phase flow approximations or account for

various operating conditions such as adverse orientations, changes in sink temperature

and ambient temperature, and insulation.

In general, the steady-state model is based on a control volume analysis. That is, the

LHP operating temperature is calculated by solving the energy balance equation for

incoming and outgoing heat flows for each LHP component. However, since the energy

balance equations are dependent on fluid properties, system pressure drop, mass flow

rate, and heat transfer coefficients which are in turn dependent on the saturation

143


APPENDIX A 144

temperature, an iterative algorithm is required. Below is a list of the assumptions

required to develop the 1-D steady-state solver (Kaya and Hoang, 1999):

1. Heat is uniformly applied to the evaporator with no losses to ambient.

2. The compensation chamber and the evaporator core contain both liquid and

vapour phases (two-phase fluid) and are therefore always at saturated conditions.

3. Single-phase flow correlations are employed to calculate the pressure drop

throughout entire LHP.

4. Heat exchange between the LHP and ambient is assumed to be due to natural

convection only.

5. The LHP achieves steady-state for a given loop condition.

As discussed in Section 2.2.1, the LHP operating temperature is established in the

compensation chamber and is a result of the energy balance equation:

QHL=Qsc+Qcc-a (A.i)

During LHP operation, the energy balance equation for the compensation chamber is

achieved only at steady-state conditions. In order to reach a steady-state condition, the

LHP self-adjusts the saturation temperature in the compensation chamber. When a

steady-state condition is achieved, this saturation temperature becomes the steady-state

LHP operating temperature (Chuang, 2003). Thus, equation 6.3 must be solved

iteratively to determine the LHP operating temperature for a given heat load.


APPENDIX A 145

A.1.1 Heat Leak

Two different flow paths exist for heat leak from the evaporator to the compensation

chamber: radially and axially. Axial heat leak is due to conduction from the evaporator

material adjacent to the compensation chamber material, and is written as:

(T - T )QHL-A=kAevap L - J ? d (A.2)

Radial heat leak is due to conduction across the primary wick (from the high pressure

side to the low pressure side) and into the evaporator core and is written as:

(~) — ^ ^ ’eff^'w ick t j , t A 'D

Since the outer and inner surfaces of the wick are at saturation conditions, the

temperature difference across the primary wick is obtained from the pressure difference

across the wick:

ATT* -

The slope of the vapour-pressure curve (£/ p)sm can be calculated by the Clausius-

Clapeyron relation. By combining radial and axial heat leak, the total heat leak is written

as:

2nk ffL ■, (T —T )Qm = , 7 \ 1 4 — Sii (A.5)

1,1 i •%:.,) 1


APPENDIX A 146

A.1.2 Effective Thermal Conductivity of the Wick

The effective thermal conductivity of the wick saturated with the working fluid is

required to calculate the radial heat leak from Eq.(A.3). Assuming a homogeneous and

isotropic wick structure, the effective thermal conductivity of the wick was obtained by

representing the volumetric average of the two phases in parallel (Kim and Golliher,

2002), yielding:

where kL is the thermal conductivity of the working fluid and s is the porosity of the

wick.

A. 1.3 Single Phase Pressure Drop

The total system pressure drop is the sum of the pressure drops of each LHP component

along the flow path. It also includes pressure drops in the wick structure, and vapour

grooves. Additionally, if the LHP is elevated against gravity, the pressure difference due

to gravity must also be taken into account. The total pressure drop across the LHP is

calculated by Eq. (2.1). The pressure drop in each section of the LHP is largely

dominated by frictional losses. For single phase duct flow of any cross-section, the

frictional pressure drop is represented by the Darcy-Weisbach equation (Holman, 1990):

k eff = K i c k { l - £ ) + £kL (A.6)

(A.7)

where:


APPENDIX A 147

64 Re 1, for laminar flow/ is the Darcy friction factor, f = \

[0.316 Re , for turbulent flow Re > 4000

Furthermore, the pressure drop across the wick is calculated using Eq. (2.3). It is again

noted that the maximum pressure created by the wick must be greater than the sum of the

pressure drops in the LHP.

A.1.4 Single Phase Heat Transfer

The overall heat transfer coefficient from the working fluid to ambient or the sink is

expressed as follows:

V L /p -a o rF -sin k1 1' — (A.8)

- +(.nNuDkL) (h0A0/

J a or sink

Eq. (A.8) assumes a constant thin wall and single phase throughout the area of heat

transfer. For fully-developed laminar flow through a circular cross-section, the Nusselt

number is a constant of either:

f 4.36 for constant surface heat flux Nud = (A. 9)

[3.66 for constant surface temperature

For fully-developed turbulent flow, the Nusselt number must be determined through

empirical correlations such as (Holman, 1990):

Nu - 0.023 Re^ Pr” (A. 10)

Where Pr is the Prandtl number is either 0.4 (for heating) or 0.3 (for cooling). Heat

exchange between the outer surface of the LHP and ambient is assumed to be through


APPENDIX A 148

natural convection. A simplified natural convection correlation for a horizontal cylinder

in air is (Holman, 1990):

h =1.32r T - T ^wall amb

D(A.11)

Heat exchange between the outer surface of the LHP and the sink may vary depending on

the cooling configuration at the condenser. Consequently, the heat transfer coefficient

between the LHP and the sink must be determined experimentally (Chuang, 2003).

A.1.5 Liquid Subcooling

The liquid temperature of single-phase fluid flow along the pipe can be calculated by

integrating the energy balance equation in the liquid-filled portions of the pipe and is

expressed as:

- m CpdTf = 'U A '

V ^ /F - a

f UA'K L J f.gink

dz (A. 12)

where, Tf is the local fluid temperature. It is noted that (U A/ j is zero along theJ \ / L ) f-sink

transportation lines since no heat is rejected to the sink. The amount of subcooling

supplied by the liquid in the liquid line is calculated using the following:

QSc = mCp{Tsat-T,lou) (A. 13)

where, Tn is the liquid line exit temperature and can be derived from Eq.(A.12).


APPENDIX A 149

A. 1.6 Loop Operating Temperature

The heat loss or gain between the compensation chamber and ambient is usually through

natural convection and is expressed as,

■r - ) (A' 14)

Finally, by substituting Eqs. (A.5), (A. 13), and (A. 14) into (A.l), the LHP operating

temperature, Tsat, can be solved iteratively given TSmk, Tamb, and Qapp.


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